Continuously variable planetary gear set

ABSTRACT

A continuously variable planetary gear set is described having a generally tubular idler, a plurality of balls distributed radially about the idler, each ball having a tiltable axis about which it rotates, a rotatable input disc positioned adjacent to the balls and in contact with each of the balls, a rotatable output disc positioned adjacent to the balls opposite the input disc and in contact with each of the balls such that each of the balls makes three-point contact with the input disc, the output disc and the idler, and a rotatable cage adapted to maintain the axial and radial position of each of the balls, wherein the axes of the balls are oriented by the axial position of the idler.

RELATED APPLICATIONS

[0001] This application claims priority to U.S. Provisional ApplicationNo. 60/494,376 filed Aug. 11, 2003, U.S. Provisional Application No.60/512,600 filed Oct. 16, 2003, U.S. Provisional Application 60/537,938filed Jan. 21, 2004 and U.S. patent application Ser. No. 10/788,736,filed Feb. 26, 2004, and all of these applications are herebyincorporated by reference in their entireties.

BACKGROUND OF THE INVENTION

[0002] 1. Field of the Invention

[0003] The field of the invention relates generally to transmissions,and more particularly the invention relates to continuously variableplanetary gear sets that can be used in transmissions as well as otherindustrial and land, air and water-borne vehicles.

[0004] 2. Description of the Related Art

[0005] In order to provide a continuously variable transmission, varioustraction roller transmissions, in which power is transmitted throughtraction rollers supported in a housing between torque input and outputdiscs, have been developed. In such transmissions, the traction rollersare mounted on support structures which, when pivoted, cause theengagement of traction rollers with the torque discs in circles ofvarying diameters depending on the desired transmission ratio.

[0006] However, the success of these traditional solutions has beenlimited. For example, in one solution, a driving hub for a vehicle witha variable adjustable transmission ratio is disclosed. This methodteaches the use of two iris plates, one on each side of the tractionrollers, to tilt the axis of rotation of each of the rollers. However,the use of iris plates can be very complicated due to the large numberof parts that are required to adjust the angular position of the irisplates during shifting of the transmission. Another difficulty with thistransmission is that it has a guide ring that is configured to bepredominantly stationary in relation to each of the rollers. Since theguide ring is stationary, shifting the axis of rotation of each of thetraction rollers is difficult.

[0007] A key limitation of this design and improvements of this designis the absence of means for generating and adequately controlling theaxial force acting as normal contact force to keep the input disc andoutput disc in sufficient frictional contact against the balls as thespeed ratio of the transmission changes. Due to the fact that rollingtraction continuously variable transmissions require various magnitudesof axial force at various torque levels and speeds in order to preventthe driving and driven rotating members from slipping on the speedchanging friction balls, where a constant level of axial force isapplied, excessive force is applied when torque transmission levels arelower. This excessive axial force lowers efficiency and causes thetransmission to fail significantly faster than if the proper amount offorce was applied for any particular gear ratio. The excessive forcealso makes it more difficult to shift the transmission. Improvements inthe field of axial force production have been made but further advancesare required.

[0008] Further improvements have been developed for the increasedperformance and efficiency of continuously variable transmissions. Thereis a need to incorporate these improvements into an advanced design fora continuously variable transmission.

SUMMARY OF INVENTION

[0009] The systems and methods illustrated and described herein haveseveral features, no single one of which is solely responsible for itsdesirable attributes. Without limiting the scope as expressed by thedescription that follows, its more prominent features will now bediscussed briefly. After considering this discussion, and particularlyafter reading the section entitled “Detailed Description of thePreferred Embodiments” one will understand how the features of thesystem and methods provide several advantages over traditional systemsand methods.

[0010] In a first embodiment a power-assisted steering system isdescribed, comprising a steering wheel, an elongated steering shaftconnected at a first end to the steering wheel and connected at a secondend to a pinion of a rack and pinion steering assembly, a motor thatprovides rotational power, a plurality of balls distributed radiallyabout the steering shaft, each ball having a tiltable axis about whichit rotates, a rotatable input disc positioned adjacent to the balls andin contact with each of the balls, a rotatable output disc positionedadjacent to the balls opposite the input disc and in contact with eachof the balls, a rotatable idler coaxial and rotatable about the steeringshaft and positioned radially inward of and in contact with each of theballs, and a tubular output shaft positioned coaxially about thesteering shaft and connected at a first end to the output disc andconnected at a second end to the pinion. In this embodiment, the axes ofthe balls are collectively responsive to an angular orientation of thesteering shaft and are adapted to orient the balls in order to convertthe rotational power of the motor to an output torque that istransmitted through the output disc to the output shaft in response to achange in the angular orientation of the steering shaft.

[0011] In some of these embodiments, a cage is described that is adaptedto maintain a radial and axial orientation of the balls about the idler,wherein the cage is adapted to rotate about the steering shaft. In someembodiments the input disc is fixed and does not rotate and the motor iscoupled to the cage.

[0012] An alternative embodiment is described further comprising; aplanetary gear set, which comprises a sun gear rotatable about thesteering shaft and coupled to the cage, a plurality of planet positionedabout, engaged with and each of which orbit the sun gear, wherein eachplanet gear rotates a planet shaft of its own, a ring gear thatsurrounds the planet gears and engages each planet gear at each planetgears furthest radial position from the steering shaft, and a generallyannular planet carrier which is rotatable about and coaxial with thesteering shaft and which retains and positions each of the planetshafts. In some of these alternative embodiments, the motor is connectedto the planet carrier and the planet shafts each extend from the planetcarrier and terminate at a connection point with the input disc so thatthe planet carrier rotates the planets about the sun gear and rotatesthe input disc about the steering shaft.

[0013] Some steering system embodiments comprise a tubular shifterhaving a first end that is dynamically attached to the idler, theshifter being angularly aligned with the steering shaft and a second endthat engages the output shaft and is positioned axially by the outputshaft such that any rotation of the steering shaft with respect to theoutput shaft moves the shifter axially, which in turn moves the idleraxially, and wherein the axes of the balls are controlled by the axialposition of the idler. Other alternative embodiments of the steeringsystem are also described

[0014] In another embodiment, a four wheeled vehicle steering system isdescribed that comprises four variable speed wheel transmissions, eachadapted to provide torque to one wheel, wherein each of the wheeltransmissions comprising, a longitudinal axis, a plurality of ballsdistributed radially about the longitudinal axis, each ball having atiltable axis about which it rotates, a rotatable input disc positionedadjacent to the balls and in contact with each of the balls, a rotatableoutput disc positioned adjacent to the balls opposite the input disc andin contact with each of the balls, and a rotatable idler coaxial aboutthe longitudinal axis and positioned radially inward of and in contactwith each of the balls. These embodiments also comprise a plurality oftorque supplies, one for each transmission, that are adapted to providea separate input to each wheel transmission, and a control systemadapted to independently control the axial position of each of theidlers in response to a request by an operator and thereby shift atransmission ratio of each of the wheel transmissions independently suchthat the wheels of the vehicle can turn at different rates causing thevehicle to turn.

[0015] Some alternative embodiments of the four wheel steering systemfurther comprise a planetary gear set mounted about the longitudinalaxis of each of the wheel transmissions.

[0016] In yet another embodiment, a hybrid vehicle is describedcomprising; a first source of rotational energy, a second source ofrotational energy, and a transmission adapted to accept rotationallyenergy from both the first and second sources. In many of theseembodiments the transmission comprises a longitudinal axis, a pluralityof balls distributed radially about the longitudinal axis, each ballhaving a tiltable axis about which it rotates, a rotatable input discpositioned adjacent to the balls and in contact with each of the balls,a rotatable output disc positioned adjacent to the balls opposite theinput disc and in contact with each of the balls, a rotatable idlercoaxial about the longitudinal axis and positioned radially inward ofand in contact with each of the balls, and a rotatable cage adapted tomaintain the axial and radial position of each of the balls. In suchembodiments, the first source supplies rotational energy to the cage andthe second energy source supplies rotational energy to the input disc.In some embodiments of the hybrid vehicle, the first source ofrotational energy is an internal combustion engine and the second sourceof rotational energy is an electric motor.

[0017] Some of the embodiments of the hybrid vehicle are described asfurther comprising an axial force generator adapted to generate acontact force between the input disc, the output disc, the balls and theidler that is proportional to an amount of torque to be transmitted bythe transmission. The axial force generator of some embodimentscomprises; a bearing disc coaxial with and rotatable about thelongitudinal axis having an outer diameter and an inner diameter andhaving a threaded bore formed in its inner diameter, a plurality ofperimeter ramps attached to a first side of the bearing disc near itsouter diameter, a plurality of bearings adapted to engage the pluralityof bearing disc ramps, a plurality of input disc perimeter ramps mountedon the input disc on a side opposite of the balls adapted to engage thebearings, a generally cylindrical screw coaxial with and rotatable aboutthe longitudinal axis and having male threads formed along its outersurface, which male threads are adapted to engage the threaded bore ofthe bearing disc, a plurality of central screw ramps attached to an endof the screw facing the speed adjusters, and a plurality of centralinput disc ramps affixed to the input disc and adapted to engage theplurality of central screw ramps.

[0018] In still other embodiments, a variable planetary gear set isdescribed comprising; a generally tubular idler, a plurality of ballsdistributed radially about the idler, each ball having a tiltable axisabout which it rotates, a rotatable input disc positioned adjacent tothe balls and in contact with each of the balls, a rotatable output discpositioned adjacent to the balls opposite the input disc and in contactwith each of the balls such that each of the balls makes three-pointcontact with the input disc, the output disc and the idler, and arotatable cage adapted to maintain the axial and radial position of eachof the balls. In such embodiments, the axes of the balls are oriented bythe axial position of the idler.

[0019] Some embodiments of the planetary gear set are describe such thatthe cage further comprises; an input stator support in the general shapeof a disc positioned between the balls and the input disc, an outputstator support in the general shape of a disc positioned between theballs and the output disc, and a plurality of spacers adapted to extendbetween and rigidly connect the input stator and output stator.

[0020] Some embodiments of the planetary gear set further comprise anaxial force generator adapted to provide a contact force between theinput disc, the output disc, the balls and the idler that isproportional to the amount of torque to be transferred through the gearset. In some of these embodiments, the axial force generator comprises agenerally disc-shaped thrust washer that is coaxial with the idler andis positioned near the side of the input disc facing away from the ballshaving a first side facing the input disc and having a set of thrustramps formed on the first side, a set of thrust-receiving ramps formedon the input disc facing the thrust washer, and a plurality of thrustelements located between and in contact with the thrust ramps and thethrust-receiving ramps.

BRIEF DESCRIPTION OF THE DRAWINGS

[0021]FIG. 1 is a schematic cutaway side view of an embodiment of atransmission shifted into high.

[0022]FIG. 2 is a partial cross-sectional view of the transmission takenalong line II-II of FIG. 1.

[0023]FIG. 3 is a schematic partial cutaway side view of the idler andramp sub-assembly of the transmission of FIG. 1.

[0024]FIG. 4 is a schematic perspective view of the ball sub-assembly ofthe transmission of FIG. 1.

[0025]FIG. 5 is a schematic cutaway side view of the cage sub-assemblyof the transmission of FIG. 1.

[0026]FIG. 6 is a schematic partial cutaway side view of an alternativeembodiment of the axial force generator of the transmission of FIG. 1.

[0027]FIG. 7 is a cutaway side view of the variator sub-assembly of thetransmission of FIG. 1.

[0028]FIG. 8 is a schematic cutaway side view of an alternativeembodiment of the transmission of FIG. 1 with two variators.

[0029]FIG. 9 is a partial cross-sectional view of the transmission takenalong line IX-IX of FIG. 8.

[0030]FIG. 10 is a perspective view of the iris plate of thetransmission of FIG. 8.

[0031]FIG. 11 is a schematic cross-sectional side view of an embodimentof an infinitely variable transmission utilizing one torque input andproviding two sources of torque output.

[0032]FIG. 12 is a schematic end-view of the embodiment of an infinitelyvariable transmission of FIG. 11.

[0033]FIG. 13 is a cross-sectional side view of an alternativeembodiment of a continuously variable transmission where the output discis part of a rotating hub.

[0034]FIG. 14 is a cross-sectional side view of an alternativeembodiment of a continuously variable transmission where the output discis part of a stationary hub.

[0035]FIG. 15 is a cross-sectional side view of another alternativeaxial force generator for any of the transmission embodiments describedherein.

[0036]FIG. 16a is a schematic side view of a power-assisted steeringsystem utilizing an infinitely variable transmission.

[0037]FIG. 16b is an alternative embodiment of the steering system ofFIG. 16a implementing an alternative ratio control mechanism.

[0038]FIG. 16c is another alternative embodiment of the steering systemof FIG. 16a implementing another alternative ratio control mechanism.

[0039]FIG. 17 is a schematic diagram of an embodiment of an infinitelyvariable transmission illustrating one possible kinematic configuration.

[0040]FIG. 18 is a schematic diagram of an embodiment of a transmissionfor use in a hybrid vehicle.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

[0041] Embodiments of the invention will now be described with referenceto the accompanying figures, wherein like numerals refer to likeelements throughout. The terminology used in the description presentedherein is not intended to be interpreted in any limited or restrictivemanner simply because it is being utilized in conjunction with adetailed description of certain specific embodiments of the invention.Furthermore, embodiments of the invention may include several novelfeatures, no single one of which is solely responsible for its desirableattributes or which is essential to practicing the inventions hereindescribed.

[0042] The transmissions and drives described herein are of the typethat utilizes speed adjuster balls with axes that tilt as described inU.S. Pat. Nos. 6,241,636, 6,322,475, and 6,419,608. The embodimentsdescribed in these patents and those described herein typically have twosides generally separated by a variator portion, to be described below,an input side and an output side. The driving side of the transmission,that is the side that receives the torque or the rotational force intothe transmission is termed the input side, and the driven side of thetransmission or the side that transfers the torque from the transmissionout of the transmission is termed the output side. As a general andabstract description of the operation of the ratio variation of many ofthe embodiments herein, an input disc and an output disc are in contactwith the speed adjuster balls. As the balls tilt on their axes, thepoint of rolling contact on one disc moves toward the pole or axis ofthe ball, where it contacts the ball at a circle of decreasing diameter,and the point of rolling contact on the other disc moves toward theequator of the ball, thus contacting the disc at a circle of increasingdiameter.

[0043] If the axis of the ball is tilted in the opposite direction, theinput and output discs respectively experience the converserelationship. In this manner, the ratio of rotational speed of the inputdisc to that of the output disc, or the transmission ratio, can bechanged over a wide range by simply tilting the axes of the speedadjuster balls. As an arbitrary assumption for use herein, the planeconnecting the centers of the balls will be considered to define theborder between the input side and the output side of the transmissionand similar components that are located on both the input side of theballs and the output side of the balls are generally described hereinwith the same reference numbers. As a convention often used in thefollowing description similar components located on both the input andoutput sides of the transmission generally have the suffix “a” attachedat the end of the reference number if they are located on the inputside, and the components located on the output side of the transmissiongenerally have the suffix “b” attached at the end of their respectivereference numbers.

[0044] Referring to FIGS. 1 and 2, an embodiment of a transmission 100is illustrated having a longitudinal axis 11 about which multiple speedadjusting balls 1 are radially distributed. The speed adjusting balls 1of some embodiments stay in their annular and spatial positions aboutthe longitudinal axis 11, while in other embodiments the balls 1 arefree to orbit about the longitudinal axis 11. The balls 1 are contactedon their input side by an input disc 34 and on their output side by anoutput disc 101. The input and output discs 34, 101 are annular discsextending from an inner bore near the longitudinal axis 11 on theirrespective input and output sides of the balls 1 to a radial point atwhich they each make contact with the balls 1. The input and outputdiscs 34, 101 each have a contact surface that forms the contact areabetween each disc 34 101, and the balls 1. In general, as the input disc34 rotates about the longitudinal axis 11, each portion of the contactarea of the input disc 34 rotates about the longitudinal axis 11 andsequentially contacts each of the balls 1 during each rotation. This issimilar for the output disc 101 as well.

[0045] The input disc 34 and the output disc 101 can be shaped as simplediscs or can be concave, convex, cylindrical or any other shape,depending on the configuration of the input and output desired. In oneembodiment, the input and output discs 34, 101 are spoked to make themlighter for weight sensitive applications, to allow ease of assembly byproviding one or more openings wherein access is provided through theinput or output disc 34, 101, and to allow fluid, such as lubricantand/or coolant to flow through input and output discs 34, 101. Therolling contact surfaces of the discs 34, 101 where they engage thespeed adjuster balls 1 can have a flat, concave, convex or other shapedprofile, depending on the torque and efficiency requirements of theapplication. A concave profile where the discs 34, 101 contact the balls1 decreases the amount of axial force required to prevent slippage whilea convex profile increases efficiency. In some embodiments the contactsurface of each of the input and output discs 34, 101 is a separatereplaceable component that can be easily removed and replaced. In suchembodiments, the contact surface can be a ring made of the appropriatematerial that is threaded into the rest of the input or output disc 34,101, while in other embodiments the contact surface has a flange orother attachment surface and is attached by fasteners. The variator 401embodiment shown in FIG. 7 illustrates an input disc 34 having an inputdisc 34 that utilizes a separate and detachable disc ring 3401 attachedby such a flange assembly. The disc ring 3401 shown in this figureattaches to the input disc 34 via a flange and fastener assembly 3402consisting of a disc ring flange (not separately identified) thatextends from the outer surface of the disc ring 3401, a disc flange (notseparately identified) that extends from outer surface of the input disc34, disc ring fasteners (not separately identified) that connect thedisc ring flange to the disc flange and a locating section 3403 that canbe used by some embodiments to precisely control the position of thedisc ring 3401 with respect to the input disc 34. The locating section3403 of the illustrated embodiment is made of an outward facing edgeformed on the disc ring flange and an inward facing edge formed on theinput disc flange, the two of which cooperate to accurately control theradial and axial position of the disc ring 3401 with respect to theinput disc 34. The use of the separate contact surface, such as the discring 3401 of FIG. 7, reduces cost by allowing for replacement of thecontact surface alone while it also allows for the use of less expensivematerials for the rest of the input and output discs 34, 101.

[0046] Additionally, the balls 1 all contact an idler 18 on theirrespective radially innermost point. The idler 18 is a generallycylindrical component that rests coaxially about the longitudinal axis11 and assists in maintaining the radial position of the balls 1. Withreference to the longitudinal axis 11 of many embodiments of thetransmission, the contact surfaces of the input disc 34 and the outputdisc 101 can be located generally radially outward from the rotationalaxes of the balls 1, with the idler 18 located radially inward from theballs 1, so that each ball 1 makes three-point contact with the idler18, the input disc 34, and the output disc 101. The input disc 34, theoutput disc 101, and the idler 18 can all rotate about the samelongitudinal axis 11 in many embodiments, and are described in fullerdetail below. The contact surfaces of the input disc 34, the output disc101 and the balls 1 can be made of, or treated with, any knowncompositions or can undergo any known material treatment to promoteadvantageous material performance characteristics of these components.Such materials and treatments are described more completely below.

[0047]FIG. 1 illustrates an embodiment of a continuously variabletransmission 100 that is shrouded in a case 40 which protects thetransmission 100, contains lubricant, aligns components of thetransmission 100, and absorbs forces of the transmission 100. A case cap67 can, in certain embodiments, cover the open end of the input side ofthe case 40, which opening allows for assembly of the internalcomponents of the transmission 100. The case cap 67 is generally shapedas a disc with a bore through its center, which allows for passagetherethrough of an input shaft 69 as described further below, and thathas a set of external threads at its outer diameter that thread into acorresponding set of internal threads on the inner diameter of the case40. Although in other embodiments, the case cap 67 can be fastened tothe case 40 using matching flanges or it can be held in place by a snapring and a corresponding groove in the case 40, and would therefore notneed to be threaded at its outer diameter. In embodiments utilizingfasteners to attach the case cap 67, the case cap 67 extends to thediameter of the case 40 so that case fasteners (not shown) used to boltthe case 40 to the machinery to which the transmission 100 is attachedcan be passed through corresponding holes in the case cap 67.

[0048] The case cap 67 of the embodiment illustrated in FIG. 1 has acylindrical portion extending from an area near its outer diametertoward the output side of the transmission 100 for additional support ofother components of the transmission 100. At the heart of theillustrated transmission 100 embodiment is a plurality of balls 1 thatare typically spherical in shape and are radially distributedsubstantially evenly or symmetrically about the centerline, orlongitudinal axis 11 of rotation of the transmission 100. In theillustrated embodiment, eight balls 1 are used. However, it should benoted that more or fewer balls 1 could be used depending on the use ofthe transmission 100. For example, the transmission may include 3, 4, 5,6, 7, 8, 9, 10, 11, 12, 13, 14, 15 or more balls 1. The provision formore than 3, 4, or 5 balls 1 can more widely distribute the forcesexerted on the individual balls 1 and their points of contact with othercomponents of the transmission 100 and can also reduce the forcenecessary to prevent the transmission 100 from slipping at the ball 1contact patches. Certain embodiments in applications with low torque buta high transmission ratio uses few balls 1 of relatively largerdiameters, while certain embodiments in applications with high torqueand a high transmission ratio can use more balls 1 or relatively largerdiameters. Other embodiments, in applications with high torque and a lowtransmission ratio and where high efficiency is not important, use moreballs 1 of relatively smaller diameters. Finally, certain embodiments,in applications with low torque and where high efficiency is notimportant, use few balls 1 of relatively smaller diameters.

[0049] Ball axles 3 are inserted through holes that run through thecenter of each of the balls 1 to define an axis of rotation for each ofthe balls 1. The ball axles 3 are generally elongated shafts over whichthe balls 1 rotate, and have two ends that extend out of either side ofthe hole through the balls 1. Certain embodiments have cylindricallyshaped ball axles 3, although any shape can be used. The balls 1 aremounted to freely rotate about the ball axles 3.

[0050] In certain embodiments, bearings (not separately illustrated) areutilized to reduce the friction between the outer surface of the ballaxles 3 and the surface of the bore through the corresponding ball 1.These bearings can be any type of bearings situated anywhere along thecontacting surfaces of the balls 1 and their corresponding ball axles 3,and many embodiments will maximize the life and utility of such bearingsthrough standard mechanical principles common in the design of dynamicmechanical systems. In some such embodiments, radial bearings arelocated at each end of the bore through the balls 1. These bearings canincorporate the inner surface of the bore or the outer surface of theball axles 3 as their races, or the bearings can include separate racesthat fit in appropriate cavities formed in the bore of each ball 1 andon each ball axle 3. In one embodiment, a cavity (not shown) for abearing is formed by expanding the bore through each ball 1 at least atboth ends an appropriate diameter such that a radial bearing, roller,ball or other type, can be fitted into and held within the cavity thusformed. In another embodiment, the ball axles 3 are coated with afriction reducing material such as babbit, Teflon or other suchmaterial. In yet other embodiments, combination bearing races are formedat each exit of the bore through each ball 1 and a correspondingcombination bearing race is formed at locations on the ball axles 3 thatcorrespond to the respective races of the ball 1. The combinationbearings utilized in such embodiments can be any type of combinationbearings and including the types described below.

[0051] Many embodiments also minimize the friction between the ballaxles 3 and the balls 1 by introducing lubrication in the bore of theball axles 3. The lubrication can be injected into the bore around theball axles 3 by a pressure source, or it can be drawn into the bore byforming rifling or helical grooves on the ball axles 3 themselves.Further discussion of the lubrication of the ball axles 3 is providedbelow.

[0052] In FIG. 1, the respective axes of rotation of each of the balls 1are shown tilted in a direction that puts the transmission in a highratio, wherein the output speed is greater than the input speed. If theball axles 3 are horizontal, that is parallel to the longitudinal axis11 of the transmission 100, the transmission 100 is in a 1:1 inputrotation rate to output rotation rate ratio, wherein the input andoutput rotation speeds are equal.

[0053]FIGS. 1, 3, and 4 illustrate how the axes of the balls 1 can betilted in operation to shift the transmission 100. Referring to FIGS. 3and 4, a plurality of legs 2, which in many embodiments are generallystruts, are attached to the ball axles 3 near each of the ends of theball axles 3 that extend beyond the ends of the holes bored through theballs 1. Each leg 2 extends from its point of attachment to itsrespective ball axle 3 radially inward toward the longitudinal axis 11of the transmission 100. In one embodiment, each of the legs 2 has athrough-bore that receives a respective end of one of the ball axles 3.The ball axles 3 preferably extend through the legs 2 such that theyhave an end exposed beyond each leg 2. In the illustrated embodiments,the ball axles 3 advantageously have rollers 4 coaxially and slidinglypositioned over the exposed ends of the ball axles 3. The rollers 4 aregenerally cylindrical wheels fitted over the ball axles 3 outside of andbeyond the legs 2 and rotate freely about the ball axles 3. The rollers4 can be attached to the ball axles 3 via spring clips or other suchmechanism, or they can ride freely over the ball axles 3. The rollers 4can be radial bearings for instance, where the outer races of thebearings form the wheel or rolling surface. Each of the rollers 4 ofsome embodiments fit over a roller shaft (not separately shown) that isseparate from the ball axles 3 and is attached to the leg 2 at aradially inward position as illustrated in FIG. 15, which allows theinput and output discs 34, 101 to have a smaller diameter. Asillustrated in FIGS. 1 and 5, the rollers 4 and the ends of the ballaxles 3 fit inside grooves 86 formed by or in a pair of stators 80 a, 80b.

[0054] The input and output stators 80 a, 80 b of the embodimentillustrated in FIGS. 1 and 5 are generally in the form of parallel discsannularly located about the longitudinal axis 11 of the transmission oneither side of the balls 1. The stators 80 a, 80 b of many embodimentsare comprised of stator discs 81 and stator curves 82. The input statordisc 81 a and output stator disc 81 b, respectively, are generallyannular discs of substantially uniform thickness with multiple aperturesto be discussed further below. Each input and output stator disc 81 a,81 b has a first side 88 that faces the balls 1 and a second side (notseparately shown) that faces away from the balls 1. Multiple statorcurves 82 are attached to the first side of the stator discs 81 a, 81 b.The stator curves 82 are curved surfaces attached or affixed to thestator discs 81 a, 81 b that each has a concave face 90 facing towardthe balls 1 and a convex face 91 facing away from the balls 1 andcontacting their respective stator discs 81. In some embodiments, thestator curves 82 are integral with or formed on the stator discs 81 a,81 b. The stator curves 82 of many embodiments have a substantiallyuniform thickness and have at least one aperture (not separately shown)used to align and attach the stator curves 82 to each other and to thestator discs 81. The stator curves 82 of many embodiments, or the statordiscs 81 a, 81 b where integral parts are used, include a slot 710 thataccepts a flat spacer 83, which allows further positioning and alignmentof the stator curves 82 and stator discs 81 a, 81 b. The flat spacers 83are generally flat and generally rectangular pieces of rigid materialthat extend between and interconnect the input stator 80 a and theoutput stator 80 b. The flat spacers 83 fit within the slots 710 formedin the stator curves 82. In the illustrated embodiment, the flat spacers83 are not fastened or otherwise connected to the stator curves 82;however, in some embodiments the flat spacers 83 are attached to thestator curves 82 by welding, adhesive, or fastening.

[0055] Also illustrated in FIG. 5, multiple cylindrical spacers 84, of agenerally cylindrical shape with bores at least in each end, areradially positioned inside of the flat spacers 83 and also connect andposition the stator discs 81 and stator curves 82. The bores of thecylindrical spacers 84 accept one spacer fastener 85 at each end. Thespacer fasteners 85 are designed to clamp and hold the stator discs 81a, 81 b, the stator curves 82, the flat spacers 83, and the cylindricalspacers 84 together, which collectively form a cage 89. The cage 89maintains the radial and angular positions of the balls 1 and aligns theballs 1 with respect to one another.

[0056] Still referring to FIGS. 1, 4 and 5, the rotational axes of theballs 1 are changed by moving either the input-side or output-side legs2 radially out from the axis of the transmission 100, which tilts theball axles 3. As this occurs, each roller 4 fits into and follows agroove 86, which is slightly larger than the diameter of the roller 4,and is formed by the space between each pair of adjacent stator curves82. The rollers 4 therefore roll along the surface of the sides 92, 93of the stator curves 82, a first side 92 and a second side 93 for eachstator curve 82, in order to maintain the plane of movement of the ballaxles 3 in line with the longitudinal axis 11 of the transmission 100.In many embodiments, each roller 4 rolls on a first side 92 of thestator curve 82 on the input side of the transmission 100 and on thecorresponding first side 92 of the corresponding output stator curve 82.The rollers 4 are slightly smaller in diameter than the width of thegrooves 86 formed between the stator curves 82, forming a small gapbetween the edges of the grooves 86 and the circumference of eachcorresponding roller.

[0057] Still referring to FIGS. 1, 4 and 5, if the opposing sets ofstator curves 82 on the input stator 80 a and output stator 80 b were inperfect alignment, in some embodiments the small gap between thecircumferences of the rollers 4 and the grooves 86 would allow the ballaxles to slightly tilt and become misaligned with the longitudinal axis11 of the transmission 100. This condition produces sideslip, asituation where the balls axles 3 are allowed to slightly movelaterally, which lowers overall transmission efficiency. In someembodiments, the stator curves 82 on the input and output sides of thetransmission 100 may be slightly offset from each other so that the ballaxles 3 remain parallel with the axis of the transmission 100. Anytangential force, mainly a transaxial force, the balls 1 may apply tothe ball axles 3 is absorbed by the ball axles 3, the rollers 4 and thefirst sides 92, 93 of the stator curves 82. As the transmission 100 isshifted to a lower or higher transmission ratio by changing therotational axes of the balls 1, each one of the pairs of rollers 4,located on the opposite ends of a single ball axle 3, move in oppositedirections along their respective corresponding grooves 86 by rolling upor down a respective side of the groove 86.

[0058] Referring to FIGS. 1 and 5, the cage 89 can be rigidly attachedto the case 40 with one or more case connectors 160. The case connectors160 extend generally perpendicularly from the radial outermost part ofthe flat spacers 83. The case connectors 160 can be fastened to the flatspacers 83 or can be formed integrally with the flat spacers 83. Theoutside diameter formed roughly by the outsides of the case connectors160 is substantially the same dimension as the inside diameter of thecase 40 and holes in both the case 40 and case connectors 160 providefor the use of standard or specialty fasteners, which rigidly attach thecase connectors 160 to the case 40, thus bracing and preventing the cage40 from moving. The case 40 has mounting holes for attaching the case 40to a frame or other structural body. In other embodiments, the caseconnectors 160 can be formed as part of the case 40 and provide alocation for attachment of the flat spacers 83 or other cage 89component in order to immobilize the cage 89.

[0059]FIGS. 1, 4, and 5 illustrate an embodiment including a pair ofstator wheels 30 attached to each of the legs 2 that roll on the concaveface 90 of the curved surfaces 82 along a path near the edge of thesides 92, 93. The stator wheels 30 are attached to the legs 2 generallyin the area where the ball axles 3 pass through the legs 2. The statorwheels 30 can be attached to the legs 2 with stator wheel pins 31, whichpass through a bore through the legs 2 that is generally perpendicularto the ball axles 3, or by any other attachment method. The statorwheels 30 are coaxially and slidingly mounted over the stator wheel pins31 and secured with any type of standard fasteners, such as snap ringsfor example. In some embodiments, the stator wheels 30 are radialbearings with the inner race mounted to the stator wheel pins 31 and theouter race forming the rolling surface. In certain embodiments, onestator wheel 30 is positioned on each side of a leg 2 with enoughclearance from the leg 2 to allow the stator wheels 30 to roll radiallyalong the concave faces 90, with respect to the longitudinal axis 11 ofthe transmission 100, when the transmission 100 is shifted. In certainembodiments, the concave faces 90 are shaped such that they areconcentric about a radius from the longitudinal axis 11 of thetransmission 100 formed by the center of the balls 1.

[0060] Still referring to FIGS. 1, 4, and 5, guide wheels 21 areillustrated that can be attached to the end of the legs 2 that arenearest the longitudinal axis 11 of the transmission 100. In theillustrated embodiment, the guide wheels 21 are inserted into a slotformed in the end of the legs 2. The guide wheels 21 are held in placein the slots of the legs 21 with guide wheel pins 22, or by any otherattachment method. The guide wheels 21 are coaxially and slidinglymounted over the guide wheel pins 22, which are inserted into boresformed in the legs 2 on each side of the guide wheels 21 andperpendicular to the plane of the slot. In some embodiments, the legs 2are designed to elastically deflect relatively slightly in order toallow for manufacturing tolerances of the parts of the transmission 100.The ball 1, the legs 2, the ball axle 3, the rollers 4, the statorwheels 30, the stator wheel pins 31, the guide wheels 21, and the guidewheel pins 22 collectively form the ball/leg assembly 403 seen in FIG.4.

[0061] Referring to the embodiment illustrated in FIGS. 1, 3, and 5,shifting is actuated by controlling the tension applied to a flexibleinput cable 155 a and a flexible output cable 155 b. Both the inputcable 155 a and the output cable 155 b extend through holes in the case40 and then through the first end of an input flexible cable housing 151a and an output flexible cable housing 151 b. The input flexible cablehousing 151 a and the output flexible cable housing 151 b of theillustrated embodiment are flexible elongated tubes that guide the inputcable 155 a and output cable 155 b radially inward toward thelongitudinal axis 11 then longitudinally out through holes in the statordiscs 81 a, b and then again radially inward where the second end of theinput and output flexible cable housings 151 a, b are inserted into andattach to the first end of input and output rigid cable housings 153 a,b, respectively. The input and output rigid cable housings 153 a, b, ofthe illustrated embodiment are inflexible tubes through which the cables155 a, b, pass and are guided radially inward from the second ends ofthe flexible cable housings 151 a, b and then direct the cables 155 a, blongitudinally through holes in the stator discs 81 a, b and toward asecond end of the rigid cable housings 153 a, b near the idler 18. Inmany embodiments, the cables 155 a, b are attached at their second endsto an input shift guide 13 a, and an output shift guide 13 b (describedfurther below) with conventional cable fasteners, or other suitableattachment means. As will be discussed further below, the shift guides13 a, 13 b position the idler 18 axially along the longitudinal axis 11and position the legs 3 radially, thereby changing the axes of the balls1 and the ratio of the transmission 100.

[0062] When output cable 155 b applies a tension force to the outputshift guide 13 b, input cable 155 a gives way and allows the idler 18 tomove axially toward the output side of the transmission 100 therebyshifting the transmission 100 toward low. When input cable 155 a appliesa tension force to the input shift guide 13 a, output cable 155 b givesway and allows the idler 18 to move axially toward the input side of thetransmission 100 thereby shifting the transmission 100 toward high.

[0063] Referring now to FIGS. 3, 4 and 5, the illustrated shift guides13 a, b, are each generally of the form of an annular ring with insideand outside diameters, and are shaped so as to have two sides. The firstside is a generally straight surface that dynamically contacts andaxially supports the idler 18 via two sets of idler bearings 17 a, 17 b,which are each associated with a respective shift guide 13 a, b. Thesecond side of each shift guide 13 a, b, the side facing away from theidler 18, is a cam side that can have a straight or flat radial surface14 towards the inner diameter of the shift guides 13 a, b, whichtransitions to a convex curve 97 towards the outer diameter of the shiftguides 13 a, b. At the inner diameter of the first side of the shiftguides 13 a, b a longitudinal tubular sleeve 417 a, b extends axiallytoward the opposing shift guide 13 a, b in order to mate with thetubular sleeve 417 a, b from that shift guide 13 a, b. In someembodiments the shift guides 13 a, b, have a convex curve 97 on theirrespective first sides from their inside diameter to their outsidediameter. In some embodiments, as illustrated in FIG. 3, the tubularsleeve of the input side shift guide 13 a has part of its inner diameterbored out to accept the tubular sleeve of the output shift guide 13 b.Correspondingly, a portion of the outer diameter of the tubular sleeveof the output shift guide 13 b has been removed to allow a portion ofthat tubular sleeve 417 a, b to be inserted into the tubular sleeve 417a, b of the input shift guide 13 a. This provides additional stabilityto the shift guides 13 a, b of such embodiments.

[0064] The cross-section side view of the shift guides 13 a, billustrated in FIG. 3 shows that, in this embodiment, the flat surface14 profile of the side facing away from the idler 18 is perpendicular tothe longitudinal axis 11 up to a radial point where the guide wheels 21contact the shift guides 13 a, b, if the ball axles 3 are parallel withthe longitudinal axis 11 of the transmission 100. From this point movingout toward the perimeter of the shift guides 13 a, b, the profile ofeach of the shift guides 13 a, b curves in a convex shape. In someembodiments, the convex curve 97 of a shift guide 13 a, b can be aradius or composed of multiple radii, or is shaped hyperbolically,asymptotically or otherwise in any other curved or curvilinear shape. Asthe transmission 100 is shifted toward low, the input guide wheels 21 aroll toward the longitudinal axis 11 on the flat 14 portion of shiftguide 13 a, and the output guide wheels 21 b roll on the convex curved97 portion of the shift guide 13 b away from the longitudinal axis 11.The shift guides 13 a, b, can be attached to each other by eitherthreading the tubular sleeve of the input shift guide 13 a with malethreads and the tubular sleeve of the output sleeve 13 b with femalethreads, or vice versa, and threading the shift guides 13 a, b,together. One shift guide 13 a, b, either the input or output, can alsobe pressed into the other shift guide 13 a, b. The shift guides 13 a, bcan also be attached by other methods such as glue, metal adhesive,welding or any other means.

[0065] The convex curves 97 of the two shift guides 13 a, b, act as camsurfaces, each contacting and pushing the multiple guide wheels 21. Theflat surface 14 and convex curve 97 of each shift guide 13 a, b contactsthe associated guide wheels 21 so that as the shift guides 13 a, b, moveaxially along the longitudinal axis 11, the guide wheels 21 ride alongthe shift guide 13 a, b surface 14, 97 in a generally radial directionforcing the leg 2 radially out from, or in toward, the longitudinal axis11, thereby changing the angle of the ball axle 3 and the rotationalaxis of the associated ball 1.

[0066] Referring to FIGS. 3 and 5, the idler 18 of some embodiments islocated in a trough formed between the first sides and the sleeveportions of the shift guides 13 a, b, and thus moves in unison with theshift guides 13 a, b. In certain embodiments, the idler 18 is generallytubular and of one outside diameter and has two sides, one near theinput stator 80 a, and one near the output stator 80 b. In otherembodiments, the outer diameter and inside diameters of the idler 18 canbe non-uniform and can vary or be any shape, such as ramped or curved.The idler 18 has an input and output idler bearing 17 a, b, on each endof its inside diameter. The idler bearings 17 a, 17 b provide rollingcontact between the idler 18 and the shift guides 13 a, b. The idlerbearings 17 a, 17 b are located coaxially around the sleeve portion ofthe shift guides 13 a, b at or near the junction of the radialextensions and the tubular sleeve of each shift guide 13 a, b, allowingthe idler 18 to freely rotate about the axis of the transmission 100.The idler bearings 17 a, b can be any type of radial or combinationradial-thrust bearing and many of the variations described below can beutilized.

[0067] A sleeve 19 is fit around the longitudinal axis 11 of thetransmission 100 inside the inside diameter of both of the shift guides13 a, b. The sleeve 19 is a generally tubular component that is held inoperable contact with an inside bearing race surface of each of theshift guides 13 a, b by an input sleeve bearing 172 a and an outputsleeve bearing 172 b. The sleeve bearings 172 a, b, provide for rotationof the sleeve 19 by rolling along an outer bearing race complimentary tothe races of the shift guides 13 a, b and can be any of the types ofbearings disclosed herein or known in the art. The idler 18, the idlerbearings 17 a, 17 b, the sleeve 19, the shift guides 13 a, 13 b, and thesleeve bearings 172 a, 172 b collectively form the idler assembly 402,seen in FIG. 3.

[0068] Referring to FIGS. 1, 2, and 3, the sleeve 19 of some embodimentshas its inside diameter threaded to engage an idler rod 171 that isthreaded into the sleeve 19. The idler rod 171 of the illustratedembodiment is a generally cylindrical rod that lies along thelongitudinal axis 11 of the transmission 100. In some embodiments, theidler rod 171 is threaded at least partially along its length to allowthreaded engagement with the sleeve 19. The first end of the idler rod171, which faces the output side of the transmission 100, is preferablythreaded through the sleeve 19 and extends out past the output side ofthe sleeve 19 where it extends into or beyond the inside diameter of theoutput disc 101. In such embodiments, the idler rod 171 is axiallypositioned by the sleeve 19, and therefore the idler 18, through thethreaded engagement. In other embodiments, the idler rod 171 can bemoved axially by a control mechanism (not shown) to position the sleeve19 and the idler 18 in order to control the transmission ratio of thetransmission 100. Some examples of such control mechanisms are disclosedbelow, although any axial positioning control mechanism known in the artcan be used to position the idler rod 171 and thereby control thetransmission ratio.

[0069] Referring to FIGS. 3 and 5, the limits of the axial movement ofthe shift guides 13 a, b define the shifting range of the transmission100. In some embodiments, axial movement of the shift guides 13 a, b islimited by inside faces 88 a, b, on the stator discs 81 a, b, which theshift guides 13 a, b contact. In some of these embodiments, at anextreme high transmission ratio, the input-side shift guide 13 acontacts the inside face 88 a on the input stator disc 81 a, and at anextreme low transmission ratio, the output-side shift guide 13 bcontacts the inside face 88 on the output stator disc 81 b. In manyembodiments, the curvature of the convex curves 97 of the shift guides13 a, b, is functionally dependent on the distance from the center of aball 1 to the center of the guide wheel 21, the radius of the guidewheel 21, the angle between lines formed between the two guide wheels 21and the center of the ball 1, and the angle of tilt of the ball 1 axis.

[0070] Referring to FIGS. 1 and 5, a spoked input disc 34 utilized insome embodiments instead of a solid disc, located adjacent to the stator80 a, partially encapsulates but generally does not contact the stator80 a. The input disc 34 may have two or more spokes or may be a soliddisc. The spokes in such embodiments reduce weight and aid in assemblyof the transmission 100. In other embodiments a solid disc can be used.The input disc 34 has two sides, a first side that contacts with theballs 1, and a second side that faces opposite the first side. The inputdisc 34 is generally an annular disk that fits coaxially over, andextends radially from, a set of female threads or nut 37 at an innerdiameter. As mentioned above, the input disc 34 is in rotating contactwith the balls 1 along a circumferential ramped or bearing contactsurface on a lip of the first side of the input disc 34, the side facingthe balls 1. As also mentioned above, some embodiments of the input disc34 have a set of female threads 37, or a nut 37, inserted into itsinside diameter, and the nut 37 is threaded over a screw 35, therebyengaging the input disc 34 with the screw 35.

[0071] Referring to FIGS. 1 and 3, the screw 35 is attached to androtated by a drive shaft 69. The drive shaft 69 is generally cylindricaland in some embodiments has an inner bore, a first end facing towardsthe output side, a second end facing toward the input side, and agenerally constant outer diameter. At the first end, the drive shaft 69is rigidly attached to and rotated by the torque-input device, usually agear, a sprocket, or a crankshaft from a motor. The drive shaft 69 hasaxial splines 109 extending from its second end to engage and rotate acorresponding set of splines (not separately identified) formed on theinside diameter of the screw 35. A set of central drive shaft ramps 99,which, on a first side facing the output side of the transmission 100,is generally a set of raised inclined surfaces on an annular disc thatis positioned coaxially over the drive shaft 69, has mating prongs thatmate with the splines 109 on the drive shaft 99, are rotated by thedrive shaft 69, and are capable of moving axially along the drive shaft69.

[0072] Still referring to FIGS. 1 and 3, a pin ring 195 contacts asecond side of the central drive shaft ramps 99, which faces the inputside of the transmission 100. The pin ring 195 is a rigid ring that iscoaxially positioned over the idler rod 171, is capable of axialmovement and has a transverse bore that holds an idler pin 196 intransverse alignment with the idler rod 171. The idler pin 196 is anelongated rigid rod that is slightly longer than the diameter of the pinring 195 and which is inserted through an elongated slot 173 in theidler rod 171 and extends slightly beyond the pin ring 195 at both itsfirst and second ends when it is inserted into the bore of the pin ring195. The elongated slot 173 in the idler rod 171 allows for axialmovement of the idler rod 171 to the right, as illustrated in FIG. 1,without contacting the pin 196 when the transmission 100 is shifted from1:1 toward high. However, when the transmission 100 is shifted from 1:1toward low, the side on the input end of the elongated slot 173 contactsthe pin 196, which then operably contacts the central drive shaft ramps99 via the pin ring 195. The idler rod 171 is thus operably connected tothe central drive shaft ramps 99 when the transmission is between 1:1and low so that when the idler rod 171 moves axially the central driveshaft ramps 99 also move axially in conjunction with the idler rod 171.The ramp surfaces of the central drive shaft ramps 99 can be helical,curved, linear, or any other shape, and are in operable contact with aset of corresponding central bearing disc ramps 98. The central bearingdisc ramps 98 have ramp faces that are complimentary to and oppose thecentral drive shaft ramps 99. On a first side, facing the output side ofthe transmission 100, the central bearing disc ramps 98 face the centraldrive shaft ramps 99 and are contacted and driven by the central driveshaft ramps 99.

[0073] The central bearing disc ramps 98 are rigidly attached to abearing disc 60, a generally annular disc positioned to rotate coaxiallyabout the longitudinal axis 11 of the transmission 100. The bearing disc60 has a bearing race, positioned near its perimeter on its side thatfaces away from the balls 1, which contacts a bearing disc bearing 66.The bearing disc bearing 66 is an annular thrust bearing at theperimeter of the bearing disc 60 and is positioned between the bearingdisc 60 and the input disc 34. The bearing disc bearing 66 providesaxial and radial support for the bearing disc 60 and in turn issupported by a bearing race on a case cap 67, which acts with the case40 to partially encapsulate the inner parts of the transmission 100. Insome embodiments, the bearing disc bearing 66 is a combination radialthrust bearing and can be any type of such bearing, such as thosedescribed below.

[0074] Referring to FIG. 1, the case cap 67 described above has atubular portion extending toward the output end from at or near itsperimeter and also having a bore through its center. The case cap 67, inaddition to the functions described above, absorbs axial and radialforces produced by the transmission 100, and seals the transmission 100,thereby preventing lubricant from escaping and contamination fromentering. As was mentioned above, the case cap 67 has a bearing racethat contacts the bearing disc bearing 66 near the perimeter of thebearing disc 60 that is located at the inside of the output end of thetubular extension from the case cap 67. The case cap 67 also has asecond bearing race facing the output side located near the insidediameter of its annular portion that mates with a drive shaft bearing104. The drive shaft bearing 104 can be a combination thrust and radialbearing that provides axial and radial support to the drive shaft 69,and can be any type of suitable bearing known in the art or describedherein. The drive shaft 67 has a bearing race formed on its outsidediameter facing the input side that mates with the drive shaft bearing104, which transfers the axial force produced by the screw 35 to thecase cap 67. An input bearing 105, adds support to the drive shaft 69and is coaxially positioned over the drive shaft 69 and mates with athird race on the input side of the inside diameter of the case cap 67opposite the drive shaft bearing 104. A cone nut 106, which is agenerally cylindrical threaded nut with a bearing race designed toprovide a running surface for the input bearing 105, is threaded overthe drive shaft 69 and supports the input bearing 105.

[0075] Referring to the embodiment illustrated in FIG. 1, a set ofmultiple perimeter ramps 61, generally forming a ring about thelongitudinal axis 11, is rigidly attached to the bearing disc 60. Theperimeter ramps 61 are multiple annular inclined surfaces that arepositioned radially about the longitudinal axis 11 and are positionedagainst or formed on the bearing disc 60 and face the output side of thetransmission 100. The inclined surfaces can be curved, helical, linear,or another shape and each one creates a wedge that produces an axialforce that is applied to a corresponding one of multiple ramp bearings62. The ramp bearings 62 are spherical but can be cylindrical, conical,or another geometric shape, and are housed in a bearing cage 63. Thebearing cage 63 of the illustrated embodiment is generally ring shapedwith multiple apertures that contain the individual ramp bearings 62. Aset of input disc ramps 64 is rigidly attached to, or formed as part of,the input disc 34. The input disc ramps 64 in some embodiments arecomplimentary to and face the perimeter ramps 61. In some embodiments,the input disc ramps 64 are also in the form of a bearing race thataligns and assist in centering the ramp bearings 62 radially relative tothe longitudinal axis 11. The ramp bearings 62 respond to variations intorque by rolling up or down the inclined faces of the perimeter ramps61 and the input disc ramps 64.

[0076] Referring now to FIGS. 1 and 3, an axial force generator 160 ismade up of various components that create an axial force that isgenerated and is applied to the input disc 34 to increase the normalcontact force between the input disc 34 and the balls 1, which is acomponent in the friction the input disc 34 utilizes in rotating theballs 1. The transmission 100 produces sufficient axial force so thatthe input disc 34, the balls 1, and the output disc 101 do not slip, orslip only an acceptable amount, at their contact points. As themagnitude of torque applied to the transmission 100 increases, anappropriate amount of additional axial force is required to preventslippage. Furthermore, more axial force is required to prevent slippagein low than in high or at a 1:1 speed ratio. However, providing too muchforce in high or at 1:1 can, in many instances, shorten the lifespan ofthe transmission 100, reduce efficiency, and/or necessitate largercomponents to absorb the increased axial forces. In some embodiments,the axial force generator 160 will vary the axial force applied to theballs 1 as the transmission 100 is shifted and also as torque is varied.In some embodiments, the transmission 100 accomplishes both these goals.The screw 35 is designed and configured to provide an axial force thatis separate and distinct from that produced by the perimeter ramps 61.In some embodiments, the screw 35 produces less axial force than theperimeter ramps 61, although in other versions of the transmission 100,the screw 35 is configured to produce more force than the perimeterramps 61. Upon an increase in torque, the screw 35 rotates slightlyfarther into the nut 37 to increase axial force by an amountproportional to the increase in torque. If the transmission 100 is in a1:1 ratio and the user or vehicle shifts into a lower speed, the idlerrod 171, moves axially toward the input side, along with the sleeve 19,sleeve bearings 172, shift guides 13 a, b, and idler 18. The idler rod171 contacts the central drive shaft ramps 99 through the pin 196 andpin ring 195, causing the central drive shaft ramps 99 to move axiallytoward the output side. The ramped surfaces of the central drive shaftramps 99 contact the opposing ramped surfaces of the central bearingdisc ramps 98, causing the central bearing disc ramps 98 to rotate thebearing disc 67 and engage the perimeter ramps 61 with the ramp bearings62 and the input disc ramps 64. The central drive shaft ramps 99 and thecentral bearing disc ramps 98 perform a torque splitting function,shifting some of the torque from the screw 35 to the perimeter ramps 61.This increases the percentage of transmitted torque that is directedthrough the perimeter ramps 61, and due to the fact the perimeter ramps61 are torque sensitive as described above, the amount of axial forcethat is generated increases.

[0077] Still referring to FIGS. 1 and 3, when shifting into low, theidler 18 moves axially towards the output side, and is pulled toward lowby a reaction of forces in the contact patch. The farther the idler 18moves toward low, the stronger it is pulled. This “idler pull,” whichincreases with an increase in normal force across the contact as well asshift angle, also occurs when shifting into high. The idler pull occursdue to a collection of transverse forces acting in the contact patch,the effect of which is called spin. Spin occurs at the three contactpatches, the points of contact where the balls contact the input disc34, the output disc 101, and the idler 18. The magnitude of theresultant forces from spin at the contact between the idler 18 and theballs 1 is minimal in comparison to that of the balls 1 and input andoutput discs 34, 101. Due to the minimal spin produced at the contactpatch of the idler 18 and ball 1 interface, this contact patch will beignored for the following explanation. Spin can be considered anefficiency loss in the contact patches at the input disc 34 and ball 1and also at the output disc 101 and ball 1. Spin produces a transverseforce perpendicular to the rolling direction of the balls 1 and discs34, 101. At a 1:1 ratio the transverse forces produced by spin, orcontact spin, at the input and output contact patches are equal andopposite and are essentially cancelled. There is no axial pull on theidler 18 in this condition. However, as the transmission 100 is shiftedtoward low for example, the contact patch at the input disc 34 and ball1 moves farther from the axis or pole of the ball 1. This decreases spinas well as the transverse forces that are produced perpendicular to therolling direction. Simultaneously the output disc 101 and ball 1 contactpatch moves closer to the axis or pole of the ball 1, which increasesspin and the resultant transverse force. This creates a situation wherethe transverse forces produced by spin on the input and output sides ofthe transmission 100 are not equal and because the transverse force onthe output contact is greater, the contact patch between the output disc101 and ball 1 moves closer to the axis of the ball 1. The farther thetransmission 100 is shifted into low the stronger the transverse forcesin the contacts become that are exerted on the ball 1. The transverseforces caused by spin on the ball 1 exert a force in the oppositedirection when shifting into high. The legs 2 attached to the ball axles3 transfer the pull to the shift guides 13 a, b, and because the shiftguides 13 a, b, are operably attached to the idler 18 and sleeve 19, anaxial force is transferred to the idler rod 171. As the normal forceacross the contact increases, the influence of spin increases at allratios and efficiency decreases.

[0078] Still referring to FIGS. 1 and 3, as the transmission 100 isshifted into low, the pull transferred to the idler rod 171 results inan axial force toward the left, as viewed in FIG. 1, which causes theinput torque to shift from the screw 35 to the perimeter ramps 61. Asthe transmission 100 is shifted into extreme low, the idler rod 171pulls more strongly, causing relative movement between the central driveshaft ramps 99 and the central bearing disc ramps 98 and shifts evenmore torque to the perimeter ramps 61. This reduces the torquetransmitted through the screw 35 and increases the torque transmittedthrough the perimeter ramps 61, resulting in an increase in axial force.

[0079] Referring to FIG. 6, a cutaway side view of an alternative axialforce generator 260 of the transmission 100 is disclosed. For purposesof simplicity, only the differences between the axial force generator160 previously described and the axial force generator 260 illustratedin FIG. 6 will be presented. The illustrated axial force generator 260includes one or more reversing levers 261. The reversing levers 261 aregenerally flat, irregularly shaped cam pieces each having an off-centermounted pivot hole with a first side radially inward of the pivot holeand a second side radially outside of the pivot hole. The first side ofthe reversing levers 261 each fit into the elongated slot 173 in theidler rod 171. When the transmission 100 is shifted toward low, the endof the elongated slot 173 contacts the first side of the reversinglevers 261 and the reversing levers 261 pivot on an axis produced by areversing pin 262 that is inserted into the pivot holes of the reversinglevers 261.

[0080] As the first sides are contacted by the end of the elongated slot173, the first side of each of the reversing levers 261 moves toward theoutput side of the transmission 100 and the second side of the reversinglevers 261 moves toward the input side of the transmission 100 therebyfulfilling the cam function of the reversing levers 261. By increasingand decreasing the length of the first side and second side, thereversing levers 261 can be designed to decrease the distance that theymove axially toward the input side and increase the force they produce.The reversing levers 261 can be designed in this manner to create amechanical advantage to adjust the axial force that they produce. Attheir second sides, the reversing levers 261 each contact the outputside of the central screw ramps 298 when the transmission 100 is shiftedtoward low. The reversing levers 261 are each attached to a lever ring263 by the reversing pins 262, which can be pressed or threaded intoholes in the lever ring 263 to hold the reversing levers 261 inposition. The lever ring 263 is a ring shaped device that fits around,and slides axially along, the idler rod 171 and has one or morerectangular slots cut through it to allow for insertion and positioningof the reversing levers 261.

[0081] Still referring to the embodiment illustrated in FIG. 6, a set ofcentral screw ramps 299 is rigidly attached to and can be rotated by thescrew 35. The central screw ramps 299 of this embodiment are similar tothe central screw ramps 99 illustrated in FIG. 3, in that the centralscrew ramps 299 are formed as ramps on the second side of a disc havinga first side facing the output side and a second side facing the inputside. As the transmission 100 is shifted toward low, the second side ofthe reversing levers 261 pushes against the first side of the centralscrew ramps 299. The central screw ramps 299, which are splined to thedrive shaft 69 via the above-described spline 109, are rotated by thedrive shaft 69, are capable of axial movement along the longitudinalaxis 11, and are similar to the central drive shaft ramps 99 of theprevious embodiment, except that the central screw ramps 299 face theinput side of the transmission 100 rather than the output side. Thecentral screw ramps 299 contact an opposing set of central bearing discramps 298, which are free to rotate relative to the drive shaft 69 andare similar to the central bearing disc ramps 98 illustrated in FIG. 3,except that the central bearing disc ramps 298 face the output side ofthe transmission 100 rather than the input side. As the central screwramps 299 are pushed axially by the reversing levers 261 toward thecentral bearing disc ramps 298, relative rotation of the ramp faces ofthe central screw ramps 299 and central bearing disc ramps 298 isdeveloped that causes the bearing disc 60 to rotate to a point such thatthe perimeter ramps 61 become engaged, thereby shifting torque to theperimeter ramps 61 and increasing the amount of axial force that isgenerated.

[0082] Referring now to FIGS. 7 and 8, an alternative embodiment of thetransmission 100 of FIG. 1 is disclosed. For the purposes of simplicity,only those differences between the transmission 1700 of FIG. 8 and thetransmission 100 of FIG. 1 will be explained. The transmission 100 ofFIG. 1 includes one variator. The term variator in this sense can, insome embodiments, be used to describe the components of the transmission100 that vary the input to output speed ratio. The assemblies andcomponents comprising the variator 401 of the present embodimentillustrated in FIG. 7 include the ball/leg assembly 403 of FIG. 4, theinput disc 34, the output disc 101, the idler assembly 402 of FIG. 3,and the cage 89 of FIG. 5. It should be noted that all components andassemblies of the variator 401 can change to best fit the specificapplication of the transmission 1700, and in FIG. 7 generic forms of theassemblies and components comprising the variator 401 are depicted.

[0083] The embodiment of the transmission 1700 illustrated in FIG. 8 issimilar to the transmission 100 of FIG. 1 but includes two variators401. This configuration is beneficial for applications where high torquecapacity is required in a transmission 1700 with a small diameter oroverall size. This configuration also eliminates bearings needed tosupport the bearing disc 114 and the output disc 101, thereby increasingoverall efficiency. Due to the fact that the transmission 1700 has twovariators 401, each variator 401 has an output side and the transmission1700 also has an output side. Thus there are three output sides and inthis configuration, the convention or marking of like components with an“a” and a “b” to differentiate between the input and output sides is notused. However, as illustrated in FIG. 8, the input side of thetransmission 1700 is to the right and the output is to the left.

[0084] Referring to FIGS. 8-9, a case 423 is illustrated that surroundsand encapsulates the transmission 1700. The case 423 is generallycylindrical and protects the transmission 1700 from outside elements andcontamination and additionally contains lubrication for properoperation. The case 423 is attached to an engine, frame, or other rigidbody (not shown) with standard fasteners (not shown), which fit throughcase holes 424. The case 423 is open on the input side, the side withthe case holes 424 or to the right as illustrated, to accept an inputtorque. Input torque is transmitted from an outside source to an inputshaft 425, which is a long, rigid, rod or shaft capable of transmittingtorque. The input shaft 425 transmits torque to a bearing disc 428 viasplines, keying, or other such manner. The bearing disc 428 is adisc-shaped rigid component capable of absorbing significant axialforces produced by the transmission 1700 and is similar in design to thebearing disc 60 illustrated in FIG. 1. An input shaft bearing 426 ispositioned coaxially over the input shaft 425 between a flange 429 onthe input end of the input shaft 425 and the bearing disc 428 to allow asmall amount of relative movement between the bearing disc 428 and theinput shaft 425. When the bearing disc 429 begins rotating, theperimeter ramps 61, ramp bearings 62, bearing cage 63, input disc ramps64, and input disc 34 rotate as previously described. This rotates theballs 1 in the first variator 420, the one on the input side.

[0085] Simultaneously, as the input shaft 425 rotates, a second inputdisc 431 is rotated. The second input disc 431 is rigidly attached tothe input shaft 425, and can be keyed with a backing nut, pressed overthe input shaft 425, welded, pinned, or attached by other methods. Thesecond input disc 431 is located on the output side of the transmission1700, opposite the bearing disc 428. The second input disc 431 and thebearing disc 428 absorb the considerable axial forces created by theperimeter ramps 61, ramp bearings 62, and input disc ramps 64 that actas normal forces to prevent slippage at the ball/disc contact patches aspreviously described. Any of the other axial force generating mechanismsdescribed herein or known in the art can also be utilized by this andother embodiments. The second input disc 431 is similar in shape to theinput disc 34 previously described and upon rotation of the input shaft425; it rotates the balls 1 in the second variator 422. The secondvariator 422 is generally a mirror image of the first variator 420 andis positioned farther from the input side of the transmission 1700 sothat the first variator 420 is situated between it and the input side.In alternative embodiments, the second input disc 431 can be splined tothe input shaft 425 and driven by a structure similar to or the same asthe bearing disc 428 of the first input disc 34. Such splines can bestandard splines or ball splines. Such embodiments allow preloading ofthe transmission with a resilient washer between the second input disc431 and its respective bearing disc-like structure (not separatelyillustrated) where the bearings and ramps at the second side areremoved. Such a structure is known in the art and is described in thereferences described and incorporated below.

[0086] As previously described, the balls 1 in the first variator 420rotate the output disc 430 through their rolling contact with thatcomponent. The output disc 430, although serving the same function asthe output disc 101 previously described, has two opposing contactsurfaces and contacts balls 1 on both variators 420, 422. From the crosssectional view illustrated in FIG. 8, the output disc 430 can be shapedin a shallow arch or upside down shallow “V,” the ends of which have acontact surface to contact the balls 1 of the two variators 420, 422.The output disc 430 surrounds the second variator 422 and extends towardthe output side in a generally cylindrical shape. In the illustratedembodiment, the cylindrical shape of the output disc 430 continuestoward the output side of the transmission 1700 surrounding the secondinput disc 431 after which the diameter of the output disc 430 decreasesand then again becomes a generally cylindrical shape of a smallerdiameter as it exits the case 423. To hold the output disc 430concentric and align it with the first and second input discs 34, 431,annular bearings 434, 435, may be used to radially align the output disc431. A case bearing 434 is positioned in the bore of the case 423 andover the output disc 430 and an output disc bearing 435 is positioned inthe bore of the output disc 430 and over the input shaft 425 to provideadditional support. The output disc 430 can be made of two pieces thatare connected together to form the illustrated output disc 430. Thisallows for assembly of the second variator 422 inside the cylindricalshell of the output disc 430. As illustrated in FIG. 8, this can beaccomplished by use of two annular flanges along the large diameter ofthe output disc 430. In some embodiments, the annular flanges arelocated generally midway along the large diameter of the output disc430.

[0087] Referring now to FIGS. 8-10, the ball axles 433 of thetransmission 1700 are similar to the ball axles 3 previously describedand perform the same function. In addition, the ball axles 433 serve asthe mechanism by which the balls 1 are tilted to vary the speed ratio ofthe transmission 1700. The ball axles 433 are elongated on each of theirrespective output sides and extend through the walls of the outputstators 435. The output stators 435 are similar to the output stators 80b previously described, but the multiple radial grooves 436 penetrateall the way through the walls of the output stators 435. The grooves 436of the output stators 435 continue all the way through the output stator435 walls so that a series of equally spaced radial grooves 436 extendradially from near the bore at the center of the output stator 435 tothe perimeter. The ball axles 433 have iris rollers 407 positionedcoaxially over their elongated output ends. The iris rollers 407 aregenerally cylindrical wheels that are capable of rotating over the ballaxles 433 and are designed to fit inside the grooves 411 of an irisplate 409. The iris plate 409 is an annular disc or plate with a borethrough its center that fits coaxially about the longitudinal axis 11 ofthe transmission 1700. The iris plate 409 is of a thickness that isgreater than twice the thickness of each iris roller 407 and has anumber of iris grooves 411 extending radially outward from near the boreto near the perimeter of the iris plate 409. As the iris grooves 411extend radially, their angular position changes as well, so that as theiris plate 409 is rotated angularly about the longitudinal axis 11, theiris grooves 411 provide a camming function along their respectivelengths. In other words, the grooves 411 spiral out from near the borein the center of the iris plate 409 to respective points near itsperimeter.

[0088] The iris rollers 407 are radiused along their outside diameters,or have fillets on their outer corners, so that their diameters remainunchanged inside the grooves 411 of the iris plate 409 when the ballaxles 433 are tilted. The iris plate 409 is of a thickness sufficient toallow iris rollers 407 from both variators 420, 422, to remain insidethe grooves 411 of the iris plate 433 at all shifting ratios. The irisgrooves 411 operate in traditional iris plate fashion and cause the ballaxles 433 to move radially inward or outward when the iris plate 409 isrotated. The iris plate 409 has a first side facing the first variatorand a second side facing the second variator and is coaxially positionedabout the longitudinal axis 11 of the transmission 1700 and overabutting bosses on tubular extensions extending from the two outputstators 435. The two output stators 435 can be attached to each otherwith conventional fasteners through axial holes (not illustrated) in thebosses of the output stators 435. The output stator 435 bosses have ahole through their centers and multiple holes positioned radiallyoutward from the center. In some embodiments, the bosses on the outputstators 435 form a space slightly wider than the iris plate 409 toprovide freedom of rotation for the iris plate 433 and some embodimentsutilize bearings between the bosses and the iris plate 409 to accuratelycontrol the position of the iris plate 409 between the output stators435. An iris cable 406 is attached to the first side of the iris plate409 near the outside diameter of the iris plate 409 and extendslongitudinally from the point of connection.

[0089] The iris cable 406 is routed through the output stator 435 of thefirst variator 420 in an orientation so that when it is pulled, itrotates the iris plate 409. The iris cable 406, after passing through anaperture near the perimeter of the output stator 435 is routed throughthe case 423 to the outside of the transmission 1700 where it allows forcontrol of the transmission ratio. An iris spring 408 is attached to thesecond side of the iris plate 409 near its outside diameter. The irisspring 408 is also attached to the output stator 435 of the secondvariator 422. The iris spring 408 applies a resilient force that resistsrotation of the iris plate 409 from tension applied by the iris cable406. When tension from the iris cable 406 is released, the iris spring408 returns the iris plate 409 to its at-rest position. Depending uponthe application of the transmission 1700, the iris plate 409 can beconfigured so that when the iris cable 406 is pulled the iris plate 409shifts the transmission 1700 to a higher transmission ratio, and whentension on the iris cable 406 is released the iris spring 408 shifts thetransmission 1700 to a low ratio. Alternatively, the iris plate 409 canbe configured so that when the iris cable 406 is pulled the iris plate409 shifts the transmission 1700 to a lower ratio, and when tension onthe iris cable 406 is released the iris spring 408 shifts thetransmission 1700 to a high ratio.

[0090] Referring to FIGS. 7 and 8, many embodiments of the transmission1700 having two variators 420, 422 require a high degree of accuracy inthe alignment of the additional rolling elements of the transmission1700. In some such embodiments, all of the rolling elements must bealigned with one another or efficiency will suffer and the lifespan ofthe transmission 1700 will be reduced. During assembly, the input disc34, the output disc 430, the second input disc 431, and the idlerassemblies 402 are aligned on the same longitudinal axis. Additionally,the cage 410, which in these embodiments consists of two cages 89 joinedby the output stators 435 as previously described, must also be alignedon the longitudinal axis to accurately position the ball/leg assemblies403. To accomplish this simply and accurately, all rolling elements arepositioned relative to the input shaft 425. A first input stator bearing440 and a second input stator bearing 444 are positioned in the bores ofthe input stators 440, 444 and over the input shaft 425 to help alignthe cage 410. An output stator bearing 442 positioned in the bore of theoutput stators 435 and over the input shaft 425 also aligns the cage410. A first guide bearing 441 is positioned in the bore of the firstshift guide 13 b and over the input shaft 425 and a second guide bearing443 is positioned in the bore of the second shift guide 13 b and overthe input shaft 425 to align the first and second idler assemblies 402.

[0091] Referring to FIGS. 8 and 9, the cage 410 is attached to the case423 with the previously described case connectors 383 that fit into caseslots 421. The case slots 421 are longitudinal grooves in the case 423that extend to the input side of the case 423, the side of the case 423that is open. In the illustrated embodiment, the case is mostly closedon the output side, which is not shown in FIG. 8, but is open on theinput side and has a mounting flange extending radially from theotherwise cylindrical body of the case 423 with case holes 424 formounting the case 423. During assembly, the transmission 1700 can beinserted into the case 423 where the case connecters 383 are aligned inthe case slots 421 in order to resist torque applied to the cage 410 andprevent the cage 410 from rotating. Case connector holes 412 in the case423 allow fasteners to be inserted into corresponding holes in the caseconnectors 383 to fasten the cage 410 to the case 423.

EXAMPLES

[0092] Each of the variations that will now be described may haveadvantageous characteristics for particular applications. The variationscan be modified and controlled as necessary to achieve the goals for anyparticular application. Specific embodiments will now be described andillustrated that employ some of the variations described herein and/orlisted in the Tables provided in U.S. patent application Ser. No.10/788,736, which were incorporated above by reference. FIGS. 11 and 12illustrate one embodiment of a transmission 1100 that is a variationhaving one source of torque input and that supplies two sources oftorque output. As before, only the significant differences between theembodiment illustrated in FIGS. 11 and 12 and the previously illustratedand described embodiments will be described. Furthermore, the componentsillustrated are being provided to illustrate to one of skill in the arthow to provide power paths and torque output sources that have not beenpreviously illustrated. It is fully understood that many additionalcomponents can and will be utilized for operational embodiments, howeverfor simplification of the drawing, many such components have beenomitted or are represented schematically as boxes.

[0093] Referring to FIG. 11, torque is input through a drive shaft 1169as in previously described embodiments. The drive shaft 1169 of thisembodiment is a hollow shaft having two ends and engaging on a first endwhatever prime mover is providing torque to the transmission 1100 andengaging at the second end a planet carrier 1130. The planet carrier1130 is a disc positioned coaxial with the longitudinal axis of thetransmission 1100 that interfaces at its center with the drive shaft1169 and extends radially to a radius near that of the inner side of thecase 1140 of the transmission 1100. In this embodiment, the case 1140 isstationary and is fixed to some supporting structure of the vehicle orequipment upon which it is utilized. A radial carrier bearing 1131 islocated between the inner surface of the case 1140 and the outer edge ofthe planet carrier 1130. The carrier bearing 1131 of some embodiments isa radial bearing that provides radial support to the planet carrier1130. In other embodiments, the carrier bearing 1131 is a compoundbearing providing both radial and axial support to the planet carrierpreventing cocking as well as radial or axial movement.

[0094] A plurality of planet shafts 1132 extend from the planet carrier1130 from a radial position between the center and the outer edge of theplanet carrier 1130. The planet shafts 1132 extend axially toward theoutput end of the transmission 1100 and are generally cylindrical shaftsthat connect the planet carrier 1130 to the input disc 1134 and eachform an axis about which a respective planet gear 1135 rotates. Theplanet shafts 1132 can be formed into the input side of the input disc1134 or the planet carrier 1130 or can be threaded into either the inputdisc 1134 or the planet carrier or can be attached by fasteners orotherwise. The planet gears 1135 are simple rotary gears that aresupported by and rotate about the planet shafts 1132 and manyembodiments utilize bearings between the planet gears 1135 and theplanet shafts 1132. They can have straight teeth or helical teeth,however where helical gears are used, thrust bearings are used to absorbthe axial thrust developed by the transmission of torque by the planetgears 1135.

[0095] Still referring to the embodiment illustrated in FIG. 11, theplanet gears 1135 engage at two areas along their respectivecircumferences at any one time as they rotate about their respectiveaxes. At a first circumferential position located farthest away from thelongitudinal axis of the transmission 1100, each planet gear 1135engages a ring gear 1137. The ring gear 1137 is an internal gear formedon or attached to the inner surface of the case 1140. In someembodiments, the ring gear 1137 is a set of radial teeth formed on theinner surface of the ring gear 1137 and extending radially inward suchthat the planet gears 1135 can engage with its teeth and ride along theinner surface of the ring gear 1137 as they orbit the longitudinal axisof the transmission 1100. At a circumferential point of the planet gears1135 generally opposite the radially outward most part, the ring gears1135 engage a sun gear 1120. The sun gear 1120 is a radial gear that ismounted coaxially about the longitudinal axis of the transmission 1100at the center of the planet gears 1135 and engages all of the planetgears 1135. As the planet carrier 1130 rotates the planet gears 1135about the sun gear 1120, the planet gears 1135 are rotated about theirrespective planet shafts 1132 by their engagement with the ring gear1137 and therefore both orbit the sun gear 1120 and rotate on their ownshafts as they orbit. This results in a rotational energy that istransmitted to the sun gear 1120 that is at a greater speed than thespeed input by the drive shaft 1169.

[0096] In the embodiment illustrated in FIG. 11, the drive shaft 1169also drives the input disc 1134 via the planet carrier 1130 and theplanet shafts 1132. However, the planet gears 1135 also drive the sungear 1120 so that the power from the planet carrier is distributed tothe input disc 1134 and the sun gear 1120. The sun gear 1120 is rigidlyconnected to and rotates the cage 1189 of this embodiment. The cage 1189is similar to the embodiments described above, and therefore not all ofthe components have been illustrated to simplify the drawing and improvethe understanding of this description. The cage 1189, as in otherembodiments, positions the balls 1101 about the longitudinal axis of thetransmission 1100 and because the cage 1189 of this embodiment rotatesabout its axis, it causes the balls 1101 to orbit the longitudinal axisof the transmission 1100. The input disc 1134, which is similar to thosedescribed above, provides an input torque to the balls 1101 in the samemanner as in previous embodiments. However the sun gear 1120 alsoprovides an input torque to the balls 1101 by rotating the cage 1189,which is added to the input from the input disc 1134. In thisembodiment, the output disc 1111 is rigidly fixed to the case 1140 anddoes not rotate about its axis. Therefore, the balls 1101 roll along thesurface of the output disc 1111 as they orbit the longitudinal axis ofthe transmission 1100 and rotate about their respective axes.

[0097] The balls 1101 cause the idler 1118 to rotate about its axis asin other embodiments, however in this embodiment, the idler 1118includes an idler shaft 1110 that extends out beyond the hole formed bythe inner diameter of the output disc 1111. The balls 1101 drive theidler 1118, which in turn drives the idler shaft 1110, which providesthe first torque output from the transmission 1100. As illustrated inFIG. 12, the idler shaft 1110 can be of a cross-sectional shape thatlends itself to easier coupling with devices that would take power fromthe idler shaft 1110 and in some embodiments, as illustrated, the shapeis hexagonal, although any such shape can be used. It is noted that dueto axial movement of the idler 1118 during shifting as described below,the idler shaft 1110 moves axially during shifting of the transmission1100. This means that the couple between the idler shaft 1110 and theoutput device (not shown) of this design allows for axial motion of theidler shaft 1118. This can be accomplished by allowing a slightly largeroutput device shaft such that the idler shaft 1110 is free to movewithin the output device, or by the use of a splined output idler shaft1110, such as by ball spline. Alternatively the idler 1118 can besplined to the idler shaft 1110 in order to maintain the axial positionof the idler shaft 1110.

[0098] Still referring to FIGS. 11 and 12, the cage 1189 can provide anoutput power source as well. As illustrated, the cage 1189 can beconnected on its inner diameter on the output side to a cage shaft 1190.In the illustrated embodiment, the cage shaft 1190 is formed at its endinto an output gear or spline to engage and supply power as a secondoutput source.

[0099] As illustrated in FIG. 11, various bearings can be implemented tomaintain the axial and radial position of various components in thetransmission 1100. The cage 1189 can be supported in its place by cageoutput bearings 1191, which are either radial bearings to provide radialsupport or are preferably combination bearings to maintain both axialand radial position of the cage with respect to the case 1140. The cageoutput bearings 1191 are assisted by cage input bearings 1192 which arealso radial or preferably combination radial-thrust bearings andposition the cage 1189 relative to the input disc 1134. In embodimentsutilizing an axial force generator where the input disc 1134 is subjectto slight axial movement or deformation, the cage input bearings 1192are designed to allow for such movement by any mechanism known in theindustry. One embodiment utilizes an outer bearing race that is splinedto the inner diameter of the input disc 1134, by a ball spline forexample, in order that the input disc 1134 can move axially slightlyrelative to the outer race of the cage input bearing 1192.

[0100] The shifting mechanism of the embodiment illustrated in FIG. 11is slightly varied from the embodiments illustrated previously in orderto allow for the transmission of output torque supplied by the idler1118. In this embodiment, the idler 1118 initiates the shifting by beingmoved axially upon actuation by the shift rod 1171 and in turn moves theshift guides 1113 axially causing the shifting mechanism to change theaxes of the balls 1101 as described above. The shift rod 1171 does notthread into the idler 1118 in the illustrated embodiment, however andonly contacts the idler 1118 via idler input bearings 1174 and idleroutput bearings 1173. The idler input and output bearings 1174, 1173,respectively, are combination thrust and radial bearings that positionthe idler 1118 both radially and axially along the longitudinal axis ofthe transmission 1100.

[0101] When the shift rod 1171 is moved axially toward the output end,the input idler bearing 1174 applies axial force to the idler, therebymoving the idler axially to the output end and initiating a change inthe transmission ratio. The shift rod 1171 of the illustrated embodimentextends beyond the idler 1118 through an inner diameter formed in thecenter of the sun gear 1120 and into the second end of the drive shaft1169 where it is held in radial alignment within the drive shaft 1169 byan idler end bearing 1175. The shift rod 1171 moves axially within thedrive shaft 1169 however and therefore the idler end bearing 1175 ofmany embodiments allows for this motion. As described before, many suchembodiments utilize a splined outer race that engages a mating splineformed on the inner surface of the drive shaft 1169. This splined raceallows the race to slide along the inner surface of the drive shaft 1169as the shift rod 1171 is moved axially back and forth and still providesthe radial support used to assist in radially aligning the shift rod1171. The inner bore of the sun gear 1120 can also be supported radiallywith respect to the shift rod 1171 by a bearing (not illustrated)located between the shift rod 1171 and the sun gear 1120. Again eitherthe inner or outer race could be splined to allow for the axial motionof the shift rod 1171.

[0102] When the idler 1118 of the embodiment illustrated in FIG. 11 ismoved axially to shift the transmission 1100, the idler 1118 moves theshift guides 1113. In the illustrated embodiment, the shift guides 1113are annular rings coaxially mounted about each end of the idler 1118.The illustrated shift guides 1113 are each held in radial and axialposition by an inner shift guide bearing 1117 and an outer shift guidebearing 1172. The inner and outer shift guide bearings of thisembodiment are combination bearings providing both axial and radialsupport to the shift guides 1113 in order to maintain the axial andradial alignment of the shift guides 1113 in relation to the idler 1118.Each of the shift guides 1113 can have a tubular sleeve (not shown) thatextends away from the idler 1118 so that the shift guide bearings 1117and 1172 can be further apart to provide additional support to the shiftguides 1113, as needed. The shift rod 1171 can be moved axially by anyknown mechanism for causing axial motion such as an acme threaded endacting as a lead screw or a hydraulically actuated piston or other knowmechanisms.

[0103] Referring to FIGS. 11 and 12, the paths of power through thetransmission 1100 follow to parallel and coaxial paths. Initially, powerenters the transmission 1100 via the drive shaft 1169. The power is thensplit and transmitted through the planet carrier 1130 both to the inputdisc 1134 and to the sun gear 1120 via the planet gears 1135. The latterpower path is then transmitted from the sun gear 1120 to the cage 1189and out of the transmission 1100 via the cage shaft 1189. This powerpath provides a fixed transmission ratio from the drive shaft based uponthe dimensions of the sun gear 1120 and the planet gears 1135. Thesecond power path is from the planet carrier 1130 through the planetshafts 1132 and to the input disc 1134. This power path continues fromthe input disc 1134 to the balls 1101 and from the balls 1101 to theidler shaft 1118 and out of the transmission 1100 through the idlershaft 1110. This unique arrangement allows the two power paths to betransmitted through the transmission 1100 not only in parallel paths butthrough coaxial paths. This type of power transmission allows for asmaller cross-sectional size for the same torque transmission and leadsto significant size and weight reductions and to a much simpler designcompared to other IVTs.

[0104] The embodiment illustrated in FIGS. 11 and 12, illustrates to oneof skill in the art how the idler 1118 can be used as a power output aslisted above and how to combine the planetary gear set with the CVT asdescribed above. It is expected that variations of this design can beutilized while achieving the various combinations described, and suchalternate designs cannot all be illustrated herein due to theoverwhelming number of combinations listed that are available. It isalso understood that the axial force generators provided herein can alsobe utilized with this embodiment, but for simplification these devicesare not illustrated. For embodiments utilizing one of the axial forcegenerators described herein, or another, it is expected that thecomponents of the axial force generator can be implemented between wherethe planet shafts 1132 connect to the input disc 1134, although otherarrangements can be employed as well. In such embodiments, the parallelpath is coaxial with the axis of the transmission 1100 allowing for amuch smaller transmission 1100 for the same torque transmission andthereby leading to reduced weight and space of such embodiments. FIGS.11 and 12 illustrate one combination in order to show how rotationalpower might be taken from the various components of the transmission invarious embodiments. Obviously, those of skill in the art will easilyunderstand how other configurations provided herein can be achieved byvarying the connections, and it would be unnecessarily burdensome andvoluminous to illustrate all or even more combinations for the simplepurpose of illustrating the combinations described. The embodimentsshown in FIGS. 11 and 12 can therefore be modified as necessary toproduce any of the variations listed above or below without the need fora separate non-coaxial parallel power path.

[0105] Referring now to FIG. 13, an alternative embodiment of atransmission 1300 is illustrated. In this embodiment, the output disc1311 is formed as part of the case of previous embodiments to form arotating hub shell 1340. Such an embodiment is suited well forapplications such as motorized two wheel vehicles or a bicycle. Asmentioned before, only the substantial differences between thisembodiment and the previously described embodiments will be described inorder to reduce the size of this description. In this embodiment, theinput torque is supplied to an input wheel 1370, which can be a pulleyfor a belt or a sprocket for a chain or some similar device. The inputwheel 1370 is then attached to the outside of a hollow drive shaft 1369by press fitting or splining or some other suitable method ofmaintaining angular alignment of the two rotary components. The driveshaft 1369 passes through a removable end of the hub shell 1340 calledthe end cap 1341. The end cap 1341 is generally an annularly shaped dischaving a bore through its center to allow passage of the drive shaft1369 into the inside of the transmission 1300 and having an outerdiameter that mates with the inner diameter of the hub shell 1340. Theend cap 1341 can be fastened to the hub shell 1340 or it can be threadedinto the hub shell 1340 to encapsulate the inner components of thetransmission 1300. The end cap 1341 of the illustrated embodiment has abearing surface and corresponding bearing on the inside of its outerdiameter for positioning and supporting the axial force generator 1360and has a bearing surface and corresponding bearing at its innerdiameter that provides support between the end cap 1341 and the driveshaft 1369.

[0106] The drive shaft 1369 fits over and rotates about an input axle1351, which is a hollow tube that is anchored to the vehicle frame 1315by a frame nut 1352 and that provides support for the transmission 1300.The input axle 1351 contains the shift rod 1371, which is similar to theshift rods described in previous embodiments, such as that illustratedin FIG. 1. The shift rod 1371 of this embodiment is actuated by a shiftcap 1343 threaded over the end of the input axle 1351 that extendsbeyond the vehicle frame 1315. The shift cap 1343 is a tubular cap witha set of internal threads formed on its inner surface that mate with acomplimentary set of external threads formed on the outer surface of theinput axle 1351. The end of the shift rod 1371 extends through a holeformed in the input end of the shift cap 1343 and is itself threadedallowing the shift cap 1343 to be fastened to the shift rod 1371. Byrotating the shift rod 1371 its threads, which may be acme threads orany other threads, cause it to move axially and because the shift rod1371 is fastened to the shift cap 1343, the shift rod 1371 is movedaxially as well, actuating the movement of the shift guides 1313 and theidler 1318, thereby shifting the transmission 1300. In otherembodiments, the shift rod 1371 does not rotate but contacts the shiftcap 1343 via bearings so that when the shift 1343 cap rotates, the shiftrod 1371 can remain in its same angular position while it is positionedaxially by the shift cap 1343.

[0107] Still referring to the embodiment illustrated in FIG. 13, thedrive shaft 1369 rides on and is supported by the input axle 1351 andone or more shaft support bearings 1372, which can be needle bearings orother radial support bearings. The drive shaft 1369 provides torque toan axial force generator 1360 as in previous embodiments. Any of theaxial force generators described herein can be used with thistransmission 1300, and this embodiment utilizes a screw 1335 that isdriven by the drive shaft 1369 by splining or other suitable mechanismthat distributes torque to the drive disc 1334 and to a bearing disc1360, as in any of the previous embodiments. In this embodiment, a driveseal 1322 is provided between the inner diameter of the input wheel 1370and the outer diameter of the input axle 1351 beyond the end of thedrive shaft 1369 in order to limit the amount of foreign material thatis admitted to the inside of the transmission 1300. Another seal (notshown) can be used between the case cap 1342 and the input wheel tolimit foreign particle infiltration from between the end cap 1341 andthe drive shaft 1369. The drive seal 1322 can be an o-ring seal, a lipseal or any other suitable seal. The illustrated embodiment alsoutilizes a similar cage 1389 as previously described embodimentshowever, the illustrated transmission 1300 utilizes axle bearings 1399to support the balls 1301 on their axles 1303. The axle bearings 1399can be needle bearings or other suitable bearings and reduce thefriction between the balls and their axles 1303. Any of the variousembodiments of balls and ball axles described herein or known to thoseof skill in the art can be used to reduce the friction that isdeveloped.

[0108] Still referring to the embodiment illustrated in FIG. 13, thecage 1389 and the shift rod 1371 are supported on the output side by anoutput axle 1353. The output axle 1353 is a somewhat tubular supportmember located in a bore formed in the output end of the hub shell 1340and between the cage 1389 and the output side vehicle frame 1315. Theoutput axle 1353 has a bearing race and bearing formed between its outerdiameter and the inner diameter of the hub shell 1340 to allow forrelative rotation of the two components as the output axle 1353 providessupport to the output side of the transmission 1300. The output shaft isclamped to the vehicle frame 1315 by an output support nut 1354.

[0109] As is illustrated in FIG. 13, this transmission 1300 is shiftedby applying tension to the shifting cord 1355 that is wrapped around andwhich applies rotational force to the shift cap 1343. The shift cord1355 is a tether capable of applying a tension force and is actuated bya shifter (not shown) used by the operator to shift the transmission1300. In some embodiments the shift cord 1355 is a guide wire capable ofboth pulling and pushing so that only one coaxial guide line (not shown)needs to be run to the shifter from the transmission 1300. The shiftingcord 1355 is conducted by housing stops 1316 to and from the shift capfrom the shifter used by the operator. The housing stops 1316 areextensions from the vehicle frame 1315 that guide the shifting cord 1355to the shift cap 1343. In the illustrated embodiment, the stop guides1316 are somewhat cylindrically shaped extensions having a slot formedalong their length through which the shifting cord 1355 passes and isguided. In other respects, the transmission 1300 illustrated in FIG. 13is similar to other embodiments illustrated herein.

[0110] Another transmission 1400 that is similar to the one illustratedin FIG. 13 is illustrated in FIG. 14. In this embodiment, the outputdisc 1411 is also fixed to the case 1440, however, the case 1440 isfixed and does not rotate. In this embodiment, however, similar to theembodiment illustrated in FIG. 11, the cage 1489 is free to rotaterelative to the output disc 1411 and the case 1440. This means that theoutput is again through the idler 1418. In this embodiment the idler1418 is attached to a moveable output shaft 1453 similar to thatdescribed in the embodiment of FIG. 11. The output shaft 1453 terminatesat the far end on the output side in an output spline 1454, which allowscoupling of the moveable output shaft 1453 to whatever device is beingsupplied with torque by the transmission 1400. In this embodiment,torque is supplied to the transmission 1400 via the input shaft 1472 bya chain and sprocket (not shown), by an input gear (not shown) or byother known coupling means. The torque then passes through to the inputdisc 1434 as described in the preceding embodiment. However, asdescribed, with reference to FIG. 13, the balls 1401 ride along thesurface of the output disc 1411 and transfer torque to the idler 1418.

[0111] As with the embodiment illustrated in FIG. 11, by supplying thetorque output via the idler 1418, the shift guides 1413 of thisembodiment are supported by bearings 1417 on the outer surface of theoutput shaft 1453. This transmission 1400 is shifted by moving the shiftrod 1471 axially and is actuated by an actuator 1443. The actuator canbe the shift cap of FIG. 13, or a wheel or gear controlled by anactuating motor or manually, or the actuator 1443 can be any othermechanism for axially positioning the shift rod 1471, such as one ormore hydraulic pistons. In some embodiments, the axial force generator1460 and the shifting mechanism illustrated below in FIG. 15 isutilized. Through this embodiment, a very high transmission ratio can beachieved at a very high efficiency and with very little frictionallosses when compared with other transmission types.

[0112] Referring now to FIG. 15, another alternative axial forcegenerator 1560 is illustrated. In this embodiment the screw 1535 islocated in the inner bore of the bearing disc (not shown) instead of theinput disc 1534. In this embodiment, the screw 1535 is driven directlyby the drive shaft (not shown) via splines 1575, which mate withmatching splines from the drive shaft. The screw 1535 then distributestorque to the input disc 1534 via central screw ramps 1598 and centraldisc ramps 1599 and to the bearing disc via its threads 1576 and acorresponding set of internal threads (not shown) formed on the innersurface of the bore of the bearing disc. As the screw 1535 is rotated bythe drive shaft, a set of central screw ramps 1598 that are formed onthe output end of the screw 1535 is rotated and engages and rotates acomplimentary set of central disc ramps 1599. The central disc ramps1599 are formed on a thrust washer surface formed on the input side ofthe input disc 1534 near its inner diameter, and as they are rotated bythe central screw ramps 1598, the central disc ramps 1599 begin to applytorque and axial force to the input disc 1534 from the reaction of theangled surfaces of the central ramps 1598, 1599. Additionally, therotation of the screw 1535 causes its threads 1576 to engage with thethreads of the bearing disc to begin to rotate the bearing disc.

[0113] Referring now to FIG. 15 in the illustrated embodiment, the axialforce generator 1560 is directly affected by the position of the idler1518. In this embodiment, the idler assembly has a tubular extensioncalled a pulley stand 1530 that extends from the input side thrust guide1513 and that ends near the input disc 1534 in an annular extensionspreading radially outward. A linkage assembly made up of a fixed link1516, a first link pin 1517, a short link 1512, a cam link 1514, a camlink pin 1515 and a stationary cam pin 1523 extends axially toward thescrew 1535 from the pulley stand 1530 and positions the screw 1535axially depending on the transmission ratio. The links 1516, 1512 and1514 are generally elongated struts. The fixed link 1516 extends fromthe input end of the pulley stand 1530 toward the screw 1535 and isconnected to the intermediate short link 1512 by the first link pin1517. The first link pin 1517 forms a floating pin joint between thefixed link 1516 and the short link 1512 such that the short link 1512can rotate about the first link pin 1517 as the two links 1516, 1512move axially during shifting. The short link 1512 is then connected atits other end to the cam link 1514 by a cam link pin 1515 and therebyforms a floating pin joint. The cam link 1514 is fixed axially by astationary cam pin 1523 that is fixed to the axle 1571 or anotherstationary component and forms a pin joint about which the cam link 1514rotates as the idler 1518 moves axially.

[0114] In the following description, for simplification of the drawing,the bearing disc 60, ramp bearings 62, perimeter ramps 61 and input discramps 64 of FIG. 1 are not separately illustrated, but similarcomponents can be utilized to fulfill similar functions in the presentembodiment. When the axial force generator 1560 illustrated in FIG. 15is in a high transmission ratio, the idler 1518 is located at an axialposition at its far input side and therefore the fixed link 1516 is alsolocated at its farthest axial point toward the input side. The firstlink pin 1517, the short link 1512 and the second link pin 1521 are alllocated towards the input side and therefore the cam link 1514 isoriented about the stationary cam pin 1523 such that its cam surface(not separately illustrated) is rotated away from the screw 1535. Thecam link 1514 applies cam force to the screw 1535 when it is rotatedabout its fixed stationary cam pin 1523 axis to force the screw towardthe output side when in low transmission ratios. However in lowtransmission ratios, as illustrated, the cam surface of the cam link1514 is rotated away from the screw 1535. This allows the screw 1535 tosettle at its farthest point towards the output side and results in thebearing disc rotating counter-clockwise, looking from the input sidetowards the output side, about the screw 1535 in order to maintainengagement with the screw threads 1576. As this occurs the bearing rampsare rotated counter-clockwise allowing the disc bearings (notillustrated here but similar to those previously described with respectto FIG. 1) to roll to a point between the bearing disc ramps and theramps of the input disc 1534 where the bearings provide little or noaxial force.

[0115] Meanwhile, due to the extreme position of the screw 1535 to theleft as viewed in FIG. 15, the central screw ramps 1598 are engaged withthe central disc ramps 1599 filly such that the input disc 1534 isrotated clockwise slightly to allow the axial position of the screw 1535in its farthest output side position. The rotation of the input disc1534 in this manner means that the input disc ramps have rotated in anopposite direction of the bearing disc ramps thereby amplifying theeffect of unloading the perimeter ramps and bearings. In such asituation, the majority or all of the axial force is being applied bythe central ramps 1598, 1599 and little if any axial force is generatedby the perimeter ramps.

[0116] As the idler 1518 moves toward the output side to shift to alower transmission ratio, the linkage assembly becomes extended as thefixed link 1516 moves axially away from the screw 1535, and the cam link1514 is rotated about the stationary cam pin 1523. As the cam link 1514is rotated about the cam link pin 1523, the axial motion of the fixedlink 1516 acts upon one end of the cam link 1514, while the other endmoves toward the screw 1535, thereby reversing the direction of theaxial force applied by the fixed link 1516. By adjusting the lengths ofwhere the various connections are made to the cam link 1514, the axialforce applied by the fixed link 1516 can be diminished or magnified bylever action. The cam end of the cam link 1514 applies an axial force toa thrust washer 1524 on the output side of the screw 1535. The thrustwasher 1524 engages a screw thrust bearing 1525 and a bearing race 1526to supply the resultant axial force to the screw 1535. In response, thescrew 1535 moves axially toward the input side and its threads 1576rotate the bearing disc clockwise, looking from input side to outputside, causing the perimeter ramps to rotate so that the ramp bearingsare moved along the perimeter ramps to a position where they begin todevelop axial force. At the same time, due to the axial movement of thescrew 1535 toward the input side, the central screw ramps 1598 aredisengaged from the central disc ramps 1599 and the input disc 1534rotates, relative to the screw 1535, counter-clockwise, again aiding themovement of the perimeter ramp bearings to a position to generate axialforce. Through this lever action of the linkage assembly, the axialforce generator 1560 of this embodiment efficiently distributes theaxial force and torque between the central ramps 1598, 1599 and theperimeter ramps.

[0117] Also illustrated in FIG. 15 is an alternative leg assembly tothat of FIG. 3 that allows for a reduced overall size of thetransmission. In the illustrated embodiment, the rollers 1504 arepositioned radially inward on the legs 1502 as compared to the legs 2 ofFIG. 3. Additionally, the input disc 34 and output disc (not shown)contact the balls 1 at a point closer to their axes which reduces theload on the idler 18 and enables the transmission to carry more torque.With these two modifications, the input disc 34 and output disc (notshown) of this embodiment can be reduced in total diameter to a diametersubstantially the same as the farthest opposing points on twodiametrically opposing balls 1501 of this embodiment as illustrated bythe line “O.D.”

[0118] Another feature of the embodiment illustrated in FIG. 15 is amodified shifting assembly. The rollers 1504 of this embodiment areformed as pulleys each with a concave radius 1505 at its outer edgeinstead of a convex radius. This allows the rollers 1504 to fulfilltheir function of aligning the ball axles 1503 but also allows them toact as pulleys to change the axes of the ball axles 1503 and the balls1501 in order to shift the transmission. The flexible cables 155described with respect to FIGS. 1 and 5, or similar shifting cables canbe wrapped around the rollers 1504 of one side so that when a tension isapplied, those rollers 1504 come closer together, thereby shifting thetransmission. The shifting cables (not illustrated in FIG. 15) can beguided through the cage (item 89 of FIG. 1) to the rollers 1504 by guiderollers 1551, which in the illustrated embodiment are also pulleysmounted on guide shafts 1552 to the output end of the pulley stand 1530.

[0119] In some embodiments, the guide rollers 1551 and the guide shafts1552 are designed to allow the axis of the guide rollers 1551 to pivotin order to maintain a pulley-type alignment with the rollers 1504 asthe ball axles 1503 change their angles with respect to the axis of thetransmission. In some embodiments, this can be accomplished by mountingthe guide shafts 1552 to the pulley stand 1530 with pivot joints ortrunnions, or any other known method. In this embodiment, one shiftcable can act on one set of rollers 1504 on either the input side or theoutput side of the balls 1501 and a spring (not shown) biases the ballaxles 1503 to shift in the other direction. In other embodiments, twoshifting cables are used with one on one side that draws the rollers1504 on its side radially inward and another cable on the opposite endof the balls 1501 that draws the rollers 1504 on its respective sideradially inward shifting the transmission thusly. In such an embodimenta second pulley stand 1530 or other suitable structure is formed on theoutput end of the shift guides 1513 and a corresponding set of guideshafts 1525 and guide rollers 1551 is mounted on that second pulleystand 1530. The cables (not shown) of such embodiments pass throughholes or slots (not shown) formed in the axle 1571 and out of thetransmission via the axle 1571. The cables can pass out of either orboth of the ends of the axle 1571 or they can pass out of additionalholes formed through the axle 1571 axially beyond either or both theinput disc (not shown) and the output disc (also not shown), or the hub(not shown) it the output disc is a rotating hub. The holes and or slotsthrough which the cables pass are designed to maximize the life of thecable material through the use of radiused edges and pulleys, and suchdesign elements are used in various locations of the axle andtransmission for conveyance of the cable.

[0120] Servo Control Systems

[0121] The embodiments described herein can be used in a servo controlsystem, such as, for example, in a power-assisted steering system. Thevariator and transmission can be utilized at or near its zero outputtransmission ratio to correct angular misalignments of a control shaftand the transmission's output shaft. In some steering embodiments, thecontinuously variable transmission is arranged coaxially with a steeringwheel or other rotary actuating member and a steering mechanism suchthat the continuously variable transmission reacts and corrects anangular misalignment between the output shaft of the transmission andthe steering shaft connected to the steering wheel.

[0122]FIG. 16a illustrates one embodiment of a servo control system usedas a power assisted steering system 1600. A steering wheel 1602 providesa direct input to a steering pinion 1675 of a rack and pinion steeringmechanism through a steering shaft 1610. The steering shaft providestorsional flexing as will be described later to provide shifting controlsignals for the power assisted steering system 1600. The steering system1600 includes the output of a constant speed electric motor 1620 that isconnected to the planet carrier 1603 via motor output gear 1621. Whilethe motor output gear 1621 engages in this embodiment by meshing withexternal teeth formed on the outer edge of the planet carrier 1603, themotor 1620 can provide input torque to the planet carrier 1603 by anymechanism known in the art such as, for example, pulley and sprocket.The planet carrier 1603 in this embodiment is connected to each of a setof planet gears 1606, which rotate about a plurality of shafts thatextend from the planet carrier 1603, and also to the input disc 1634.The planet gears 1606 engage at their radially outward side with thering gear 1607, which is fixed and does not rotate, and at theirradially inward side with the sun gear 1605. Therefore, the planet gears1606 rotate the sun gear 1605 at a fixed rotation rate determined by thespeed of the electric motor 1620, the radii of the planet gears 1606 andthe radius of the sun gear 1605.

[0123] The variator 1640 of this embodiment acts as a variable planetarygear set in series with the fixed planetary gear set made up of the ringgear 1607, the planet gears 1606 and the sun gear 1605. The sun gear1605 drives the cage 1689 of the variator 1640 and the planet carrier1603 drives the input disc 1634. The torque provided to the cage 1689and the torque provided to the input disc 1634 are summed by thevariator 1640 and transmitted to the output disc 1601. The output disc1601 drives a power assist shaft 1615 in this embodiment, which adds theadditional torque to assist the manual steering applied to the steeringwheel 1601 in this embodiment.

[0124] In other embodiments, the motor 1620 provides input torque to thesun gear 1605 directly, which drives the planet gears 1606 and theplanet carrier 1603, thereby driving the input disc 1634. In suchembodiments, the rotational speed transmitted to the input gear 1634,and therefore the balls (not separately referenced in this figure) andthe idler 1618 is significantly reduced.

[0125] In still other embodiments of the steering system 1600, theplanetary gear set is removed. The motor 1620 of this embodimentprovides input directly to the cage 1689 and the input disc 1634 isfixed to the case (not separately identified). In such embodiments, theconstruction and design of the steering system 1600 is simplified.

[0126] Still referring to FIG. 16a, the amount and rotational directionof the output assisting torque applied by the power assist shaft 1615 isdetermined by the transmission ratio of the variator 1640. The system1600 accomplishes the control of the transmission ratio through amechanical connection of the variator 1640 and the steering wheel 1602.As a driver turns the steering wheel 1601 to turn the steering wheels ofa vehicle, the steering shaft 1610 is rotated and thereby begins torotate the steering pinion 1675. In a typical steering system, thesteering pinion 1675 engages a steering rack (not shown) that is typicalof a rack and pinion steering system. The rack is connected at each endvia steering tie rods (not shown) to steering arms on the hubs of thesteering wheels of the vehicle (all not shown). Such components arestandard items in steering systems.

[0127] As the driver begins to apply torque to the steering shaft 1610by turning the steering wheel, the steering shaft 1610 transmits thattorque to the steering pinion 1675, which engages the rack to convertthe rotational torque of the steering wheel 1601 into linear motion ofthe ends of the rack, which is then transferred to the wheels via thetie rods and steering arms. This applies a moment to the wheel thattends to rotate each wheel about its turning axis of rotation, which isresisted by the frictional contact of the tire and the road. As the roadresists the turning of the tire, the torque applied to the steeringshaft 1610 must be increased to cause the wheels to turn. The steeringshaft is designed with a flexural modulus that allows the steering shaftto begin to torsionally flex at a desired torque level in response tothe torque applied to the steering wheel 1602. Because the power assistshaft 1615 is attached to the steering pinion 1675 coaxially with thesteering shaft 1610, as the steering shaft 1610 begins to torsionallyflex, as just described, it becomes angularly misaligned with the powerassist shaft 1615. This angular misalignment is used in this embodimentto shift the variator 1640.

[0128] In the embodiment illustrated in FIG. 16a, a generally tubularshifter 1632 is angularly aligned with and axially moveable along thesteering shaft 1610. The shifter 1632 is a relatively short tube that issplined to, or otherwise angularly aligned with, a portion of thesteering shaft 1610. The shifter 1632 has a first end near the idler1618 and a second end facing away from the idler 1618. The first end ofthe shifter 1632 is designed to dynamically connect to the idler 1618 sothat as the idler 1618 rotates during operation of the variator 1640,the shifter 1632 can move the idler 1618 axially in order to shift thetransmission ratio of the variator 1640. In the illustrated embodiment,the first end of the shifter 1632 has a flange extending radiallyoutward from the rest of the tubular body of the shifter. The first endof this shifter 1632 fits within a recess 1619 of the idler 1618 and isheld within the idler 1618 by a retention ring 1621. The retention ring1621 can be a snap ring 1621 or can have a threaded outer diameter toscrew into the recess 1619. Thrust bearings 1617 allow the idler 1618 torotate relative to the shifter 1632 while allowing the shifter 1632 toapply an axial force to move the idler 1618 axially, in order to shiftthe idler 1618.

[0129] Still referring to FIG. 16a, a steering pin 1630 extends in atransverse manner through the shifter 1632 and fits into a spiralingslot formed in the power assist shaft 1615. The steering pin 1630 slidesalong the spiral slot in the power assist shaft 1615. As the steeringshaft 1610 begins to torsionally flex, the shifter 1632 and the steeringpin 1630 begin to become angularly misaligned with the power assistshaft 1615. The spiral shape of the slot in the power assist shaft 1615causes a camming effect that moves the shifter 1632 axially, dependingon the direction of the angular misalignment. The axial movement of theshifter 1632 drives the idler 1618 to move axially and shift the powersteering system 1600 to a transmission ratio that produces an output inthe output disc 1601, which output acts to correct the angularmisalignment. As the angular misalignment is corrected, as the vehicleattains the appropriate turning attitude, the steering pin 1630, theshifter 1632 and the idler 1618 begin to ease back to their respectivezero-output positions and the output disc 1601 provides less or nooutput torque.

[0130] When the power assist shaft is applying no torque, such as when avehicle is traveling straight, the variator 1640 is at a ratio providingzero output. When the power assist shaft is applying some power assistin the clockwise direction as viewed in FIG. 16a along the steeringshaft 1615 from right to left, the variator 1640 is in a ratio providinga slight output torque in that direction. When the power assist shaft1615 is applying some power assist in the counter-clockwise direction asviewed in FIG. 16a along the steering shaft 1610 from right to left, thevariator 1640 is in a ratio providing a slight output torque in thatdirection. Therefore, the entire range of the ratios available inseveral embodiments of FIG. 16a will be around the zero output range ashigh ratios are not typically required in such applications. However, inother applications, higher ratios ranges may be necessary and themechanical attachments should be designed to optimize the ratio rangefor each application.

[0131] Still referring to FIG. 16a, this system 1600 provides a way ofadjusting the output of the variator 1640 to respond directly andmechanically to the action of the driver in turning the steering wheel.This system 1600 is merely one example of a directly responsive shiftingmechanism and many other such control mechanisms can be used. The keycharacteristics of many embodiments of such a control circuit for asteering system is that as the steering wheel 1602 begins to turn, thevariator 1640 should begin to apply an output rotation in theappropriate direction until an equilibrium is reached between the torqueapplied to the steering wheel 1602, the power assist provided by thepower assist shaft 1615 and the feedback force provide by the wheels ofthe vehicle, or rudder if in a water borne vessel, through the rack andsteering pinion 1675.

[0132] The steering system 1600 illustrated in FIG. 16a accomplishes itscontrol function mechanically, but this can easily be programmed into ahydraulic or electric control system, as described herein, to achievesimilar or different results as desired. FIG. 16b illustrates anadditional embodiment utilizing an alternative shifting mechanism. Inthis embodiment, the axial position of the idler 1618 is controlled in amanner similar to that described above for FIG. 16a, in that the idler1618 has a recess 1619 in one of its ends, the input end in this case,and the shifter 1632 now extends from within the recess 1619 in theidler 1688 towards the steering wheel 1602. The shifter 1632 has aflange 1633 in this embodiment that is retained within the recess 1619by thrust bearings 1617 and a retention ring 1621 similar to theanalogous or same components of the previously described embodiments.The shifter 1632 extends toward the steering wheel 1602 and extendsbeyond the planet carrier 1603 where it terminates having a lead screw1660 formed on this end.

[0133] Still referring to FIG. 16b, the lead screw 1660 can be any setof threads formed on the outer surface of the shifter 1632, but someembodiments utilize acme threads. The lead screw 1660 is engaged by aset of internal threads 1661, which is a set of complimentary threadsfacing inward and engaging with the lead screw 1660 and which aremounted on the inside of a shift ring 1664. The shift ring 1664 is atubular ring having the internal threads 1661 formed upon its internalsurface and which is rotated about the steering shaft angularly by ashifting motor 1662. The shifting motor 1662 is mounted to a fixedsurface and is capable of rotating the shift ring 1664 about thesteering shaft in either direction in order to engage the internalthreads 1661 with the lead screw 1660 and thereby move the shifter 1632axially depending on the direction of rotation of the shift ring 1664.The shifting motor 1662 is an electric motor in the illustratedembodiment, but could also be hydraulic or pneumatic. In alternativeembodiments, the lead screw 1660 is replaced by a piston (not shown) andthe shift ring 1664 and internal threads 1661 are replaced with apneumatic or hydraulic cylinder (not shown), wherein the piston ispositioned within the cylinder by a hydraulic or pneumatic controlcircuit, which are common in the art.

[0134] Still referring to FIG. 16b, the activation of the shifting motor1660 is determined by an indicator 1665 and a sensor 1666. The indicator1665 is mounted on the steering shaft 1610 and indicates any angularmotion by the steering shaft 1610 to the detector 1666. The detector1666 is arranged radially around the indicator 1665 and detects themagnitude and direction of the angular rotation of the steering shaft1610. The indicator 1665 and detector 1666 can be any type of componentcapable of fulfilling their described functions such as, but not limitedto, rotary encoders or any other such devices. The indicator 1665 anddetector 1666 can also comprise multiple components such as where theindicator 1665 is an annulus extending from the steering shaft 1610 andthe detector 1666 is capable of reading the position or motion of theannulus. In one embodiment, the indicator 1665 comprises an annular gearthat moves a rack, which creates linear displacement and the detector1666 is a linear encoder capable of very fine motion detection.

[0135] Still referring to FIG. 16b, the detector 1666 provides a motionsignal to a controller 1667 via one or more signal lines 1668 and thecontroller 1667 sends motor control signals to the motor via one or morecontrol lines 1669. The controller 1666 can sample the position ormotion of the steering shaft 1610 by controlling the detector 1666, orthe detector 1666 can provide a set of position signals to thecontroller 1667 at a specific rate. The faster the motion signals aresampled, the more sensitive the response of the power assistance. Forexample, in one embodiment the controller can receive signals from thedetector at a rate of between 5 and 20 million signals per secondalthough higher or lower frequencies can be used as are common in theindustry. In other embodiments, the indicator 1665, the detector 1666,the controller 1667, the signal lines 1668 and the control lines 1669are replaced by the power steering pump and rotary valve system ofcurrent power steering systems in conjunction with the cylinder andpiston described above. Such a system is common in current steeringsystems and can be implemented as described herein.

[0136]FIG. 16c illustrates yet another alternative embodiment for apower assisted steering system 1600. Only the differences between theembodiments illustrated in FIGS. 16 b and c will be discussed. In theillustrated embodiment, the shifter 1632 is angularly aligned with thesteering shaft 1610 by splines 1663 or keyways and keys or other suchstructure and thereby rotates with the steering shaft 1610. In someembodiments, ball splines or other low friction structures are used asthe splines 1663 in order to ease the turning force required by theoperator. As with the previous embodiment, the lead screw 1660 and theinternal threads 1661 engage one another to create axial movement of theshifter 1632 in reaction to rotation of the steering wheel 1602.However, in this embodiment, the internal screws 1661 are fixed by aretaining ring 1662 to a support structure rather than to a rotatingmotor. Therefore, the internal threads 1661 do not rotate about thesteering shaft 6 10.

[0137] In the illustrated embodiment, as the steering wheel 1602 isrotated by an operator, the splines 1663 rotate the shifter 1632, whichrotates the lead screw 1660, which engages with the internal threads1661 to develop an axial force that changes the axial position of theshifter 1632 in order to change the position of the idler 1618 anddevelop an output torque to respond to the steering of the operator. Thegain or reaction rate of the steering system 1600 response of theillustrated embodiment to the input steering by the operator can beadjusted by controlling the pitch of the internal threads 1661 and thecorresponding lead screw 1660. The shifting mechanisms described for thevarious embodiments illustrated in FIGS. 16b-c can be used for any ofthe transmission embodiments described or incorporated herein in orderto achieve advantageous shifting control and manipulation.

[0138] Gearing Systems

[0139] Due to the extremely configurable nature of the embodiments ofthe IVTs and CVTs described and incorporated herein, and the high degreewith which the components can be easily scaled in size to accommodatethe amount of torque and rotational power to be transmitted, the IVTsand CVTs make extraordinarily advantageous gear sets. Either reductiongears or step-up gears can be configured by the various embodimentsherein as the input disc, output disc, and variator of each embodimentcreate a continuously variable planetary gear set as described herein.The addition of an additional fixed ratio planetary gear set oradditional CVTs lined up in successive arrangement allows designers toachieve infinite gearing ratios and flexibility. For instance, thevariator 401 of FIG. 7 can be combined with a planetary gear set asillustrated in FIGS. 11, 16a-c to create the speed reduction systemillustrated in FIG. 17. FIG. 17 is a schematic view of a transmissionsystem 1700 that can suffice as such a continuously or infinitelyvariable gear set. The illustrated transmission system 1700 includes aplanetary gear set 1730, a variator 1740, and an output shaft 1710 thatreceive and transmit rotational energy from an input torque source 1720.The torque source 1720 can be an engine, a motor, a piece of industrialequipment, a differential, a shaft, or any other source of rotationalenergy. Additionally, although this schematic illustration shows theoutput shaft 1710 connected to the output disc 1711 of the variator1740, it should be understood that the output disc 1711 can easily beconnected to the cage 1789 or the idler 1718 of the variator 1740 aswell, as described and illustrated herein and in copending U.S. patentapplication Publication Ser. No. 10/788,736 incorporated above(hereinafter “the '736 application), and this description applies tothose embodiments as well. Furthermore, although the illustratedembodiment shows the ring gear 1737 as being fixed and the planetcarrier 1733 and the sun gear 1735 of the planetary gear set 1730 asbeing attached to the input disc 1734 and the cage 1789 of the variator1740, respectively, any of the connective combinations identified in thetables of the '736 application can be used and the following discussionapplies to those as well.

[0140] Such variable gear sets can be used effectively in any systemthat utilizes or transmits rotational energy or converts linear motioninto rotational motion or vice versa. In systems where a variable inputspeed is provided and a fixed or relatively constant output speed isdesired, the embodiments of the CVTs or IVTs described herein areexceedingly advantageous and useful. For instance, superchargers andturbochargers for combustion engines have efficiencies and performancecharacteristics that vary as a function of rotational speed eitherindependently of one another or even dependent upon one another.However, the prime movers for these components, direct connection to thecrankshaft for the supercharger and a turbine driven by exhaust gasesfor the turbocharger, also vary in supply speed or force depending onthe rotation speed of the engine, which varies with throttle position.Therefore, in such applications, a variable speed gear set such as thosedescribed herein can be used to reduce the adverse effects of thechanging input speed when a desired output speed of the pump of thesecomponents is desired.

[0141] For instance, a supercharger is typically utilized ondiesel-powered vehicles such as semi-tractor trailers used in long-haultransportation of goods. The boost in intake pressure supplied from thesupercharger to the engine is affected by the rate of rotation of thecrankshaft of the engine. It is desired to maintain the speed ofrotation of the supercharger near a target speed at various conditions.Existing superchargers use a fixed ratio speed changing gear set tochange the engine speed to the rough speed range used by thesupercharger. Through the use of a continuously variable gear set asdescribed herein, the speed of the supercharger could stay in a smallerrange of operational speeds over the entire range of engine speedsthereby allowing increased efficiency of the supercharger over theentire range of engine speeds. This is also true for any pump or turbineapplication. Most, if not all, centrifugal pumps and turbines haveperformance characteristics that vary with the speed of their respectiveprime movers. In all such applications, the use of the continuouslyvariable gear sets described herein can be used to maintain theperformance of these items in their preferred operational windows evenas the speed of their respective prime movers varies within or out ofthe resultant desired speed range.

[0142] The planetary gear set 1730 can be any ordinary planetary gearset or it can be any advancement in such structures. For example, U.S.patent application Publication No. 2003/0232692 (hereinafter “the '692application”), the entire disclosure of which is hereby incorporated byreference, discloses an example of an advance in planetary gear systemsthat can be implemented with the embodiments of IVTs disclosed herein.The variability created by the advancement disclosed in the '692application can be utilized to further increase the variability of theIVTs described herein, allowing such embodiments to fulfill even morefunctions. Similarly, U.S. patent application Publication No.2003/0153427, the entire disclosure of which is hereby incorporated forall that it discloses, discloses an advancement in planetary gearsystems in which the planetary gear set varies the input to output speedratios as a function of the load transmitted through the system. Again,such a system can be used in the IVTs described herein to create agreater range of effective ratios, or to vary the torque transferringcapabilities of the transmission system.

[0143] Control Mechanisms and Systems and Protocols

[0144] Many advances have been made for controlling the transmissionratio of past CVT designs such as toroidal and adjustable pulley CVTs.Many of these control systems can be adjusted and revised to takeadvantage of the advanced design and increased efficiency of the IVTsand CVTs described herein. For instance, U.S. patent applicationPublication No. 2003/0228953 A1 (hereinafter the '953 application)describes a control system and shifting protocol for a CVT that isutilized on a variable pulley-type CVT that can be adjusted as describedherein to take advantage of many of the embodiments described above toprovide a shifting control protocol and system, and the entiredisclosure of that application is incorporated herein by reference. Inthe CVT described as being controlled in that published application,clutches and brakes are required that allow the transmission of thatapplication to shift from forward transmission ratios to reversetransmission ratios. Many of the embodiments described herein allow atransmission to shift from its highest forward transmission ratio to itshighest reverse transmission ratio through its continuous shifting rangewithout changing the engagement of any of the components of thetransmission.

[0145] The '736 application incorporated above describes numerouscombinations of input, throughput and output of the embodiments of theIVTs described therein. Many of those can be successfully utilized asthe transmission for a vehicle such as a car. In one embodiment, such asthat illustrated in FIG. 17, where the planetary gear set 1730 ispositioned on the input side of the variator 1740, the crank shaft 1720from the engine of a car is provided as input to the planet carrier 1753of the planetary gear set 1730 of the transmission 1700, the cage 1789is free to rotate, the ring gear 1737 is fixed to the case (notseparately identified) of the transmission 1700 or to some fixed supportstructure of the vehicle, the idler 1718 is free to rotate and theoutput shaft 1710 is connected to the output disc 1711. In such anembodiment, the range of possible transmission ratios is affected by theratio of the circumference of the ring gear 1737 to that of the planetgears 1736, or the PG ratio for this configuration. Some embodimentsutilize a PG ratio between 1.5 and 10. In other embodiments, the PGratio is between 2 and 5, while in still other embodiments a PG ratio ofbetween 3 and 4 is utilized. Some embodiments use a PG ratio of 3.

[0146] Some embodiments of the IVTs described herein conforming to thesePG ratios provide transmission ratios adequate for many applications andprovide high efficiency, suitable transmission ratio range andoperational simplicity for nearly any vehicle using such a transmission.For instance, with a PG ratio of 3, some IVTs of the embodimentsdescribed herein that are configured as just described can provideengine input to transmission output ratios ranging from about 2.5forward to 0 forward all the way to 2.5 in reverse without everdisengaging any of their components. This setup also allows the coaxialalignment of the input shaft and the output shaft, thereby leading todecreased overall size, reduced and simplified resultant torsionalstresses and various other advantages known to those of skill in theart.

[0147] To incorporate the control functions of the embodiments describedin the above-mentioned '953 application with many of the IVTs describedand incorporated herein, the clutches and brakes are removed and the ECUdescribed in that control system is operably connected to the servocontrols or the pneumatic or hydraulic controls utilized to control theIVTs so that the control system can be implemented and its advantagescan be exploited and amplified. The fully continuous, manually shiftedand staged protocols described in the '953 are all employed with many ofthe IVTs herein to provide performance that is significantly improvedover the embodiments described in the '953 application.

[0148] Another example of the use of embodiments of the IVTs describedherein as an advantageous improvement of existing technology can beillustrated with reference to U.S. patent application Publication No.2003/0109347 A1 (hereinafter “the '347 application”), the disclosure ofwhich is incorporated herein in its entirety. In the '347 application, ahydromechanical IVT is utilized on a tractor to maximize thefunctionality of the tractor where multiple speeds are desirable forvarious functions. Again, the IVT of that embodiment utilizes clutchesand brakes to vary its speed over its range of transmission ratios. TheIVT described in the '347 application, as with others like it, utilizesa parallel power path, that is two paths through which rotational poweris transmitted from the input to the output that are not collinear withone another. This configuration requires a synchronization of componentsof the transmission in order to shift the various stages and realize thefull transmission ratio range. This adds unnecessary parts andcomplexity, and therefore cost, to the transmission. In contrast, manyof the IVTs described and incorporated herein utilize a collinear pairof power paths that do not require synchronization or clutching andbraking in order to vary the transmission ratio over its entire range.Additionally, because power can be output via any one or many possiblecombinations of the output disc, the cage and the idler, the IVTsdescribed herein, provide for both power output to the drive train aswell as a power takeoff unit so that the same transmission can performboth functions simultaneously.

[0149] Referring also to U.S. patent application Publication No.2001/0044358 (hereinafter “the '358 application) the entire disclosureof which is incorporated herein, another system for controlling a CVT isdescribed that responds to requests by vehicle operators for changes invehicle performance. In this embodiment, only belt-and-pulley andtoroidal CVTs are contemplated, which require parallel power paths aswell as synchronizing of components and the use of brakes and/orclutches to shift modes throughout the range of transmission ratiosincluding reverse transmission ratios. Many embodiments of IVTs and CVTsdescribed and incorporated herein can be advantageously implementedalong with the CVT control system of the '358 application, as well asother publications incorporated above and below, by removing the controlsystem and functions that require manipulation or adjustment of theforward/backward switching mechanism. Specifically, the manipulation ofthese components adds an additional calculation in the response to ademand for a change in driving conditions made by the driver. Throughthe use of the certain embodiments of the IVTs described andincorporated herein, such as for example the embodiment illustrated inFIG. 17, the response is simpler for the electronic control unit toemploy and there is less chance of failure and a smoother resultingspeed variation over the entire range of driving conditions.

[0150] Furthermore, because many of the IVTs and CVTs described andincorporated herein are analogous if not similar to existing planetarygear set-based automatic transmissions, many of the existing advancesfor controlling existing automatic transmissions can be advantageouslyemployed on those IVTs, while employing the CVT and IVT controlprotocols described in the incorporated patents and publishedapplications. For example, U.S. patent application Publication No.2003/0027687, the entire disclosure of which is hereby incorporated byreference, discloses a control system that operates the engine inconjunction with a transmission controller. Any of the transmissioncontrol systems described herein can be used with such a control systemin order to maximize vehicle efficiency regardless of enginedisplacement. Other such improvements can advantageously be employed aswell.

[0151] As a further example, many of the embodiments of IVTs and CVTsdisclosed and incorporated herein can also be advantageously employed inconjunction with the control systems disclosed in U.S. patentapplication Publication Nos. 2003/0162633 (hereinafter “the '633application”), 2003/0158646 (hereinafter “the '646 application”), and2001/0046924 (hereinafter “the '924 application”), the disclosures ofall of which are hereby incorporated in their entireties. While the '633application and the '646 application both operate a CVT that appears tolack reversing functions on its own, and the '924 application operates aCVT that includes the reversing mechanisms of other past advances, whichinclude a planetary gear set and clutches and brakes, all of theseapplications control a belt-and-pulley CVT that is hydraulically orpneumatically controlled. Therefore, each of these applications requirethe manipulation and control of brakes and clutches in order to achievethe complete transmission ratio range spanning from high forward to highreverse. This means that the power train throughout the transmissionundergoes connections and disconnections as the transmission ratio isvaried over the entire range of ratios, and this may lead to decreasedperformance, safety or component life. The present embodiments of theIVTs and CVTs that utilize these control systems achieve their functionsthroughout their transmission ratio ranges without the switching andbraking previously required.

[0152] U.S. Pat. No. 6,390,946 (hereinafter “the '946 patent”), theentire disclosure of which is hereby incorporated by reference,discloses a system designed to assist in the sensing of rotationalspeeds of various components. The '946 patent discloses the constructionof a sensing system that can be applied to any of the rotatingcomponents of the IVTs and CVTs described and incorporated herein inorder to provide speed signals to the transmission control system.Additionally, U.S. patent application Publication Nos. 2002/0095992 and2003/0216216, the entire disclosures of both of which are herebyincorporated by reference, both describe additional sensing points andsystems of a rolling traction CVT that can be utilized by the controlunits of embodiments described and incorporated herein to optimize theperformance of the engine and transmission of those embodiments.

[0153] The signals provided by such sensing systems can be utilized bythe systems described above or by U.S. patent application PublicationNos. 2002/0173895, 2003/0135316, 2003/0135315, 2003/0045395,2003/0149520 and 2003/0045394, the entire disclosures of all of whichare hereby incorporated by reference. These are additional controlsystems that can be implemented for use with the IVTs and CVTs describedand incorporated herein. As mentioned previously, only thebelt-and-pulley and toroidal CVTs were contemplated for use with thesecontrol systems and therefore the functional components and commandscontrolling the forward/reverse switching brakes and clutches can beremoved to allow control of the present embodiments. Furthermore,whether the method of shifting any particular embodiment is electricmotor, pneumatic or hydraulic piston or any other method, the systemsincorporated herein can be adapted to such shifting mechanisms by anymethod known to those of skill in the art in order to achieve theadvantages of the present IVTs and CVTs as controlled by the controlsystems described and incorporated above and below.

[0154] Furthermore, many advances have been made in the specific area ofhydraulic control systems for controlling toroidal and belt-and-pulleytype CVTs. Many of these systems and advances can be implemented for usein the hydraulically controlled embodiments of the IVTs and CVTsdescribed and incorporated herein. For example, U.S. Pat. Nos.5,052,236, 5,090,951, 5,099,710, 5,242,337, 5,308,298, 6,030,310,6,077,185, 6,626,793 and 6,409,625 as well as U.S. patent applicationPublication Nos. 2003/0158009, 2003/0114259, 2003/0228952, 2002/0155918,2002/0086759, 2002/0132698 and 2003/0158011, the entire disclosures ofall of which are hereby incorporated by reference, disclose hydrauliccontrol systems and control fluid systems as well as pressure system foruse in either a toroidal or a belt-and-pulley transmission system. Thesecontrol systems and circuits can be implemented on the IVTs and CVTsdescribed and incorporated herein by adapting these systems to operatethe piston of the hydraulically shifted transmission systems describedherein. Furthermore, U.S. Pat. No. 6,464,614 discloses a hydraulicsystem that provides hydraulic supply circuitry or passages in thecasing containing the remainder of the transmission system. Any or allof these systems or advances, or even combinations of them, arebeneficial in various applications of the IVT and CVT embodimentsdescribed and incorporated herein.

[0155] Such hydraulic control systems can include feedback controlinformation as well. U.S. patent application Publication Nos.2003/0050149, 2002/0169051, 2002/0155910, the entire disclosures of allof which are hereby incorporated by reference, each discloses ahydraulic control system for an existing CVT or IVT. These publicationsalso disclose the monitoring of certain system parameters to be feddirectly back into the control circuit, either mechanically orelectronically, to adapt the controls to the response of thetransmission system to the existing control signal. Such feedbacksignals can provide very advantageous effects when utilized along withthe control systems described above for use with the IVTs and CVTsdescribed and incorporated herein, such as preventing hunting for theproper output speed, reducing overall time to achieve the desired speedchange, and increased overall vehicle efficiency.

[0156] However, these applications describe control units that areutilized on toroidal or belt-and-pulley CVTs but that can beadvantageously employed with many of the IVTs and CVTs described herein.Again, by removing the switching of clutches and brakes that must beaccomplished in the past transmissions, all of the advantages disclosedin these published applications can be enhanced. The hydraulic controlsthat operate the sheeves or pulleys of these transmissions can besimplified to operate the hydraulic piston and cylinder control systemused to control certain embodiments of the IVTs and CVTs as describedabove. Furthermore, the circuitry, controls and the functional signalsthat manipulate the clutches and brakes of these three publishedapplications can be removed and replaced with a control regime thatsimply adjusts the ratio of the IVT or CVT throughout its entire range.Many of the IVTs and CVTs described herein also allow removal of thetorque converter of the '924 application and any clutches that may beutilized with that advance. However, these components can still beutilized in certain embodiments as conditions may dictate.

[0157] For example, some embodiments utilize a clutch prior to thetransmission that controls an amount of torque applied to thetransmission, independent of the variability of the torque supplied bythe engine. In many of such embodiments, control systems are utilizedthat adjust the clutch in order to prevent slippage of the rollingcontact surface. U.S. patent application Publication No. 2003/0069682,the entire disclosure of which is hereby incorporated by reference,discloses a control system and protocol that is used by such embodimentsto control and prevent slippage of the clutch and the transmission.

[0158] Control Protocols

[0159] In addition to these and other systems that can control a CVT oran IVT, there are many control protocols that can be utilized tomaximize the advantages of such a transmission in a vehicle. Because ofthe inherent differences, and indeed advantages, of a CVT or an IVT ascompared to a standard geared transmission, operational paradigms can beabandoned in order to achieve the increased efficiency and performanceavailable from these advanced designs. Several advances have been madein the area of control protocols for CVTs or IVTs that can beimplemented for use with many of the embodiments of the IVTs and CVTsdescribed and incorporated herein. U.S. Pat. No. 5,820,513 (hereinafter“the '513 patent”), U.S. patent application Publication Nos.2003/0229437 (hereinafter “the '437 application”) and 2003/0022752(hereinafter “the '752 application”) relating to establishingoperational protocols for controlling a CVT as engine speed varies, andU.S. patent application Publication No. 2002/0028722 (hereinafter “the'722 application”) relating to a control system and protocol for anexisting IVT, each disclose ways of controlling CVTs or IVTs, and theentire disclosures of all these publications are hereby incorporated byreference. These publications each describe systems and methods foroperating a variator of a CVT that is used in a vehicle to optimize theperformance of the vehicle; however, these publications only contemplatethe use of existing toroidal or belt-and-pulley transmissions andtherefore would benefit greatly from the implementation of manyembodiments of the IVTs and CVTs described and incorporated herein.

[0160] There are other examples of control protocols and performancemapping methods that have been developed for existing CVTs and IVTs aswell. For Example U.S. patent application Publication Nos. 2003/0119630relating to mapping of CVT performance and function to develop shiftingstrategies, 2002/0165063 relating to controlling the emissions andtreating the intake of the engine in conjunction with transmissioncontrols for increased efficiency and/or performance, 2002/0062186relating to operating a CVT while traveling uphill or downhill,2002/0082758 relating to calculating a target speed ratio andcontrolling the CVT according to the target, the entire disclosures ofall of which are hereby incorporated by reference, each disclosecontrolling methods and techniques that are employed for use withcertain embodiments of the IVTs and CVTs described and incorporatedherein.

[0161] Other examples of such control protocols that can be employedalong with some of the IVT and CVT embodiments described herein areprovided in U.S. patent application Publication Nos. 2002/0132697relating to controlling a CVT having a multi-stage torque sensor,2003/0022753 relating to simultaneous controlling of a CVT and an enginein response to a requirement for power output, 2003/0060681 relating tospecific control equations for operating a toroidal CVT, 2003/0119627relating to determining the transmission ratio of a CVT, 2003/0004030relating to specific methods for operating a CVT for increasingefficiency and performance, 2002/0128115 and 2002/0115529 relating toestablishing a target speed ratio and generating a creep torque basedupon the difference between the target speed ratio and the actual speedratio, and 2002/0072441 relating to compulsory down-shifting of thetransmission based upon various conditions, the entire disclosures ofall of which are hereby incorporated by reference. The embodimentsutilizing one or more of these advances achieve various functionalityand performance advantages that make these embodiments desirable forvarious applications. However, again, because these publications onlycontemplate either the toroidal or belt-and-pulley transmissions, theyinclude control componentry and functionality to control theforward/reverse mechanisms and can be optimized for use with theembodiments described herein by removal of such components andfunctionality.

[0162] Again, these publications describe inventions that are improvedthrough the benefits of the simpler and more versatile design of many ofthe present embodiments including the variator 1740 and transmissionsystem 1700 described above and illustrated in FIG. 17. The collinearmultiple power paths of the IVTs described herein provide not -onlysmaller designs, with simpler torsional reactive forces, but as notedbefore, also allow shifting of the transmission throughout its entirerange of ratios without the need for mode shifting brakes and clutches.In certain embodiments, this results in the input being connected to theoutput in the same manner over the entire range of transmission ratios,thereby leading to increased component life and performance as well assimpler bearing wear and other advantages. As a contrast, the IVT systemof the '722 application requires the actuation of a recirculation clutchor a direct clutch depending on the particular driving conditionsdemanded by the driver. Furthermore, the enclosing case of many of thepresent embodiments described herein can be made in a much simplermanner as fewer bearing and support surfaces need to be incorporatedinto the case. In some embodiments of the present IVTs and CVTs thatimplement the control systems and protocols disclosed in the '513patent, the '437 application, the '752 application and the '722application, the control mechanisms and functions that actuate theseclutches or brakes are removed from the control routines in order tosimplify the control system and protocols. By doing so, theseembodiments allow a simpler system and protocol for operating andcontrolling the IVTs or CVTs while still realizing all of the advantagesof such transmissions.

[0163] Alternative Architecture

[0164] Certain embodiments also take advantage of other mechanicaladvances in transmission systems. For example, as stated above, multipleplanetary gear sets can be combined to form compound systems of gearingto operate in unison with the CVT of certain embodiments of the IVT inorder to add additional range or functionality to the resultingtransmission system. U.S. patent application PublicationNo.2002/0169048, the entire disclosure of which is hereby incorporatedby reference, discloses compound planetary gear sets in order tofacilitate the use of a toroidal CVT in an IVT, however, thispublication also suggests how multiple planetary gear sets may bealigned or combined in order to functionally combine them. Throughreference to the illustrations and accompanying descriptions of thatpublication, present IVT or gearing embodiments can be created thatutilize such compound gearing.

[0165] Furthermore, U.S. patent application Publication No.2003/0125153, the entire disclosure of which is hereby incorporated byreference, discloses a vehicle having power transmitted from the enginevia a CVT to all four of its wheels. Embodiments of the IVTs or CVTsdescribed herein are easily incorporated advantageously for use on sucha vehicle power train for improved performance and to reduce maintenanceassociated with the transmission. Certain embodiments of IVTs and CVTsdescribed and incorporated herein incorporate advances disclosed in U.S.patent application Publication No. 2003/0186769 relating to twoplanetary gear sets coupled to each other and to two variators in acompound arrangement, and the entire disclosure of that publication ishereby incorporated by reference. This compounding and the various modesavailable are good examples of how such combinations can be effectivelyincorporated in certain embodiments of the IVTs and CVTs describedherein. U.S. patent application Publication No. 2003/0220167, the entiredisclosure of which is also incorporated herein by reference, disclosesa CVT that employs multiple sets of planet gears in its planetary gearset. Embodiments of the IVTs and CVTs described and incorporated hereinutilizing multiple sets of planetary gears for additional transmissionrange and additional advantages benefit from the disclosure of thispublication in carrying out such compounding.

[0166] Also, advances in planetary gears themselves are exploited bycertain embodiments. For instance, certain embodiments utilize advancessuch as that described in U.S. patent application Publication No.2003/0171183, the entire disclosure of which is hereby incorporated byreference. This publication discloses a speed ratio amplifier for usewith advanced CVT and CVT control systems to amplify the speed changingeffect of a planetary gear set. Embodiments of the IVTs and CVTsdescribed and incorporated herein can achieve even greater ratio rangesfor the transmission as a whole.

[0167] Related Technology

[0168] Additional technological advances can be implemented for use inthe IVTs and CVTs described and incorporated herein as well. Forinstance, because the embodiments described herein are rolling tractionforms of transmissions, lubrication is required for many embodiments andadvances in the field of lubrication can be advantageously implementedto promote the proper and efficient operation of those embodiments. Forexample, the methods and systems of lubrication described in U.S. patentapplication Publication No. 2002/0183210 and U.S. Pat. No. 6,500,088 areemployed in many embodiments to advantageously lubricate thetransmissions of those embodiments, and the entire disclosure of both ofthose publications are hereby incorporated by reference. Additionally,the lube oils disclosed in U.S. patent application Publication No.2003/0013619 are used in many embodiments as traction and lubricatingoils, and the entire disclosure of that application is herebyincorporated by reference.

[0169] The lubricating systems of many embodiments, as well as thetransmission components themselves can require additional heatdissipation. U.S. Pat. No. 5,230,258, which is hereby incorporated byreference in its entirety, discloses a method of providing cooling tothe transmission. Certain embodiments utilize a casing that utilizes thecooling channels described therein in order to provide the proper amountof cooling to the lube oil and transmission components.

[0170] Furthermore, improvements that have been made for toroidal CVTsthat relate to generation and control of axial traction force areemployed in some embodiments. For example, the biasing mechanismdescribed in U.S. Pat. No. 4,893,517, the entire disclosure of which ishereby incorporated by reference, is utilized in some embodiments of theCVTs where it replaces the more complex axial force generators (“AFGs”)described herein and in some embodiments of the IVTs where it can easilybe positioned on the output side of the variator or between the planetcarrier and the input disc. The planet carrier of such embodiments candrive the cam flange of the AFG and the input disc is modifiedaccordingly to accept the thrust and the torque from the cam flange. Theimprovements to such AFGs that are described in U.S. Pat. Nos. 6,287,235and 6,514,171, the entire disclosures of both of which are herebyincorporated by reference, are utilized by some embodiments utilizingsuch AFGs.

[0171] Additionally, the double-sided preloading described in U.S. Pat.No. 4,968,289 (hereinafter “the '289 patent”), the entire disclosure ofwhich is hereby incorporated by reference, is utilized in someembodiments where the preloading spring of that publication ispositioned on the output side of the IVT or CVT embodiment or ispositioned on the same side as the cam flange. For instance, in someembodiments the preloading springs are located between the planetcarrier and the case and apply force to the planet carrier that istransmitted to the input disc, while in other embodiments, the springsare located between the case and the second input disc in the dualcavity design, which applies a force against the input shaft asillustrated in the '289 patent. Many embodiments utilize the integraltorque sensor disclosed in U.S. patent application Publication No.2002/0111248, the entire disclosure of which is hereby incorporated byreference, as an input to control systems to adjust the axial force, forhydraulic and pneumatic AFG embodiments, or as an input for slipdetection functions or for any other function.

[0172] A hydraulic AFG is utilized in some embodiments to carefullycontrol the axial force applied according to the torque to betransmitted. U.S. patent application No. 2003/0100400 (hereinafter “the'400 application”), the entire disclosure of which is herebyincorporated by reference, discloses a hydraulic AFG. Some embodimentsimplement this design by creating a two part output disc having a pistonpart facing the ball and a cylinder part that houses the piston part andthereby creates a chamber between the two parts that is sealeddynamically, as illustrated in the '400 application. As pressure isapplied to the chamber, the two parts tend to separate and the pistonpart is pressed against the balls. Other embodiments implement this byattaching the planet carrier to the cylinder part and forming the inputdisc as the piston part. In such embodiments, the axial force can becarefully planned over the range of torques to be applied, and moreimportantly, can be adjusted or corrected without changing thecomponents of the AFG. U.S. patent application No. 2003/0109340, theentire disclosure of which is hereby incorporated by reference,discloses a dynamic seal that many embodiments implementing hydraulicAFGs utilize to improve their respective performances.

[0173] As another example of the implementation of advances made fortoroidal CVTs, certain embodiments herein utilize the taper bearingsdescribed in U.S. Pat. No. 5,984,827 to act as combination bearings.Several combination thrust-radial bearings are described for use in theembodiments of the CVTs and IVTs described and incorporated herein, andmany if not all such bearings can benefit through the implementation ofadvances in such bearing technology. Furthermore, some embodimentsutilize one or more of the improvements to these AFGs disclosed in U.S.Pat. No. 5,027,669 related to implementation of an axially moveableshaft, U.S. Pat. No, 5,899,827 related to a loading cam design, U.S.patent application Publication No. 2003/0017907 related to lubricationof ball splines, U.S. Pat. No. 5,984,826 relating to retaining thebiasing mechanism, U.S. patent application Publication No. 2002/0111244disclosing a hydraulic AFG, U.S. patent application Publication No.2003/0078133 related to a preloader accompanied by a hydraulic AFG andU.S. Pat. No. 5,027,668 related to creating a centrifugal lubricationreservoir at the AFG, all of which are hereby incorporated by referencein their respective entireties. U.S. Pat. No. 6,248,039, the entiredisclosure of which is hereby incorporated by reference, discloses animprovement to the use of ball splines for the mounting of a disc to ashaft where the disc and the shaft can move axially with respect to oneanother. Some embodiments utilize this improvement for at least one oftheir splines, regardless of the AFG in use.

[0174] U.S. Pat. No. 6,312,356, the entire disclosure of which is herebyincorporated by reference, discloses a way to accommodate a certainamount of flexing of the input or output disc. Some embodiments utilizesuch an improvement on at least one of the input or output discs toaccommodate a certain amount of elastic deformation of that tractiondisc, or those discs. U.S. Pat. No. 5,267,920 (hereinafter “the '920patent”) discloses the use of pilot holes to angularly align componentsduring manufacture, and its entire disclosure is incorporated herein byreference. Certain embodiments utilize pilot aligning holes as describedin the '920 patent in order to correctly align the components of any orall of the variator, the AFG or any other components.

[0175] Several advances have been made in the treatment and preparationof materials for use in rolling traction CVTs and IVTs that are utilizedby certain embodiments as well. For instance, some of the bearings ofsome embodiments experience high load and/or high cycling and thereforebenefit from bearing advances made for other mechanical applications.Some of the bearings that can experience high load and/or high cyclingare the ball axle bearings, the idler support bearings, and othersimilar bearings. For instance, some embodiments described hereinutilize for one or more of their bearings, bearings made according toU.S. patent application Publication No. 2003/0219178, which isincorporated herein by reference in its entirety. Additionally, therolling elements of some or all bearings of some embodiments arecontained in bearing races formed according to U.S. patent applicationNo. 2002/0068659 the entire disclosure of which is hereby incorporatedby reference. Such bearing races can improve performance of the bearingover the life of the component.

[0176] In some embodiments, bearings that are expected to experiencehigh levels of stress are treated as disclosed in U.S. patentapplication Publication No. 2002/0082133, the entire disclosure of whichis hereby incorporated by reference. In some embodiments, at least apart of one or more of the input disc, output disc, balls, idler, or anyof the high-stress bearings of the IVT or CVT is, or are, manufacturedas described in any or all of U.S. patent application Publication Nos.2002/0086767, 2003/0087723, 2003/0040401, 2002/0119858 and 2003/0013574,the entire disclosures of all of which are hereby incorporated byreference. In embodiments where the rolling contact surfaces of theinput and output discs are detachable, the rolling surfaces are treatedfor hardness as disclosed by these publications while the input discs ofsome such embodiments are manufactured for strength and durability asdisclosed. Furthermore, the bearing cages that retain many of thebearings of some embodiments are manufactured according to U.S. patentapplication Publication No. 2002/0151407, the entire disclosure of whichis hereby incorporated by reference.

[0177] In addition to these material composition and treatment advances,some embodiments utilize technology that is specific to the field ofrolling traction transmissions. For instance, the traction surface ofeither or both of the input and output discs disclosed in U.S. Pat. No.6,527,667, the entire disclosure of which is hereby incorporated byreference, vary in roughness. Some embodiments herein apply suchvariation to at least one of the input disc, the output disc and theballs so that at certain ratios the active surface will have a differentsurface roughness than that for at least one different ratio. Similarly,at least one of the traction surfaces of some embodiments conforms tothe disclosure of U.S. Pat. No. 6,524,212, which is hereby incorporatedby reference in its entirety, to control and improve the traction oilfilm thickness.

[0178] In another manufacturing advancement, U.S. patent applicationPublication No. 2003/0096672, which is hereby incorporated by referencein its entirety, discloses the use of a datum on the output disc bywhich the rest of the disc is manufactured. In some embodiments, thisconcept has been incorporated and a radially flat surface is provided onthe input and/or output disc that acts as an indexing origin for themanufacture and fitting of the rest of the disc(s). In some embodiments,this flat surface occurs near the inside bore.

[0179] U.S. Pat. No. 6,159,126, which is hereby incorporated byreference in its entirety, discloses a method of preventing a shock of aCVT where a vehicle's engine may start while the transmission hasdrifted away from the lowest ratio. Some embodiments utilize a biasingmechanism in order to mechanically return the transmission to a zerooutput or other desired orientation, according to the incorporatedpatent to prevent such a shock from occurring. In some embodimentsutilizing hydraulic control systems, this is accomplished by a spring ofappropriate biasing direction and force for the particular application.

[0180] Additional Applications

[0181] Many embodiments of the IVTs and CVTs described herein areadvantageously implemented in various applications such as agricultural,aerospace, aircraft, watercraft, industrial machinery and auto racingamong others. Certain advances have been made that utilize existing CVTtechnology that would see increased performance that could not have beencontemplated by the original inventors when those existing CVTs werereplaced by the IVTs and CVTs of many embodiments, described herein. Forexample, U.S. Pat. No. 4,922,788 (hereinafter “the '788 patent”), theentire disclosure of which is hereby incorporated by reference,discloses the use of two IVTs for use on a twin track-driven vehicle,one IVT for each track. By changing the output rotation speed of eachIVT independently, the operator can steer the vehicle without need forturning wheels or other steering system. The IVTs operate independentlyof one another to provide either forward or reverse rotation to theirrespective tracks to drive the vehicle. The existing IVTs utilized inthe 788 patent suffer all of the same defects as described above, namelythe toroidal CVT is inherently unstable and the ratio control system isalso inherently unstable, and requires in any practical embodiment aparallel power path and clutches and brakes to perform its IVT function.Due to the inherently unstable design, the toroidal CVT requiressignificant structural strength for its support and to house its controlsystem.

[0182] Therefore, the embodiments described herein provide smaller andsimpler components that reduce cost, size maintenance and increasereliability. The embodiments herein allow use of CVTs and IVTs not onlyon heavy two track vehicles, but also on the wheels of tractors andlight tractor equipment. A vehicle can be provided with a relativelysmall and lightweight transmission at every wheel to have all-wheelsteering where the steering is provided by the transmission ratio ofeach particular transmission.

[0183] In another application, some embodiments of the CVTs and IVTsdescribed and incorporated herein are used in place of the existing CVTsdisclosed for use in U.S. patent application No. 2002/0165060(hereinafter “the '060 application”), the disclosure of which is herebyincorporated by reference in its entirety. The torque distributionsystem described in the '060 application is greatly enhanced by thecomparatively smaller CVTs and IVTs described herein and because theresulting input and output axes of such embodiments are collinear. Suchan orientation makes embodiments of the present application an idealcandidate for use in the torque distribution system of the '060application and indeed makes such a system even more practicallyfeasible.

[0184] Another advantageous application of embodiments described hereinis a hybrid vehicle, which is a vehicle with two power sources, asillustrated in FIG. 18. FIG. 18 is a schematic diagram of an embodimentthat could be used for a hybrid vehicle without using a planetary gearset. For example, the combustion engine 1820 of a gas-electric hybridcan provide the input into the cage 1889 while the electric motor 1820provides torque input to the input disc 1834. The variator 1840 sums thetorque of the two inputs and provides a resulting output to the outputshaft 1810. In an alternative embodiment, inputs are switched so thatthe combustion engine 1820 of the gas-electric hybrid provides the inputinto the input disc 1834 while the electric motor 1820 provides torqueinput to the cage 1889. In other alternative embodiments, a planetarygear set is adapted to the input side in a similar manner as that ofFIG. 17 and torque is provided by the internal combustion engine 1820directly to the cage 1889 through a bore (not shown) in the sun gear andthe electric motor provides torque to the planet carrier, or vice versa.Such embodiments allow greater flexibility in designing a system thatoptimizes the efficiency of both torque sources, but add complicationand cost to the transmission 1800.

[0185] U.S. patent application Publication No. 2003/0032515 (hereinafter“the '515 application”) discloses a system for use in a gas-electrichybrid vehicle, and its entire disclosure is hereby incorporated byreference. However, the '515 application requires two electricalmachines to operate, at any one time one acting as a motor and the otheracting as a generator. Embodiments of the IVTs and CVTs described hereinare utilized in a vehicle as described in the '515 application, allowingremoval of the variable gear ratio by the engine and the second machine.Therefore, this leads to a much simpler design.

[0186] The embodiments described herein are examples provided to meetthe descriptive requirements of the law and to provide examples. Theembodiments described herein are examples provided in order to explainand to facilitate the full comprehension and enablement of all that isdisclosed herein and the description of these examples is not intendedto be limiting in any manner. Therefore, the invention is intended to bedefined by the claims that follow and not by any of the examples orterms used herein. Additionally, terms utilized herein have been used intheir broad respective senses unless otherwise stated. Therefore, termsshould not be read as being used in any restrictive sense or as beingredefined unless expressly stated as such.

What is claimed is:
 1. A power-assisted steering system, comprising: anelongated steering shaft having a first end and a second end andconnected at the second end to a pinion of a rack and pinion steeringassembly; a motor that provides rotational power; a plurality of ballsdistributed radially about the steering shaft, each ball having atiltable axis about which it rotates; a rotatable input disc positionedadjacent to the balls and in contact with each of the balls; a rotatableoutput disc positioned adjacent to the balls opposite the input disc andin contact with each of the balls; a rotatable idler coaxial androtatable about the steering shaft and positioned radially inward of andin contact with each of the balls; and a tubular output shaft positionedcoaxially about the steering shaft and connected at a first end to theoutput disc and connected at a second end to the pinion; wherein theaxes of the balls are collectively responsive to an angular orientationof the steering shaft and are adapted to orient the balls in order toconvert the rotational power of the motor to an output torque that istransmitted through the output disc to the output shaft in response to achange in the angular orientation of the steering shaft.
 2. The steeringsystem of claim 1, further comprising: a cage adapted to maintain aradial and axial orientation of the balls about the idler, wherein thecage is adapted to rotate about the steering shaft.
 3. The steeringsystem of claim 2, wherein the input disc is fixed and does not rotateand wherein the motor is coupled to the cage.
 4. The steering system ofclaim 2, further comprising a planetary gear set, which comprises; a sungear rotatable about the steering shaft and coupled to the cage; aplurality of planet gears positioned about, engaged with and each ofwhich orbit the sun gear, wherein each planet gear rotates about aplanet shaft of its own; a ring gear that surrounds the planet gears andengages each planet gear at each planet gears furthest radial positionfrom the steering shaft; and a generally annular planet carrier which isrotatable about and coaxial with the steering shaft and which retainsand positions each of the planet shafts; wherein the motor is connectedto the planet carrier and wherein the planet shafts each extend from theplanet carrier and terminate at a connection point with the input discso that the planet carrier rotates the planets about the sun gear androtates the input disc about the steering shaft.
 5. The steering systemof claim 2, further comprising a tubular shifter having a first end thatis dynamically attached to the idler, the shifter being angularlyaligned with the steering shaft and a second end that engages the outputshaft and is positioned axially by the output shaft such that anyrotation of the steering shaft with respect to the output shaft movesthe shifter axially, which in turn moves the idler axially, and whereinthe axes of the balls are controlled by the axial position of the idler.6. The steering system of claim 2, further comprising a tubular shifterhaving a first end dynamically attached to the idler and a second endfacing away from the idler and having a lead screw formed thereon,wherein the shifter is angularly aligned with the steering shaft,wherein the lead screw interacts with a set of internal threads toaxially position the idler in response to any rotation of the steeringshaft.
 7. The steering system of claim 2, further comprising: a tubularshifter positioned coaxially about the steering shaft and having a firstend that is dynamically attached to the idler and a second end thatfaces toward the steering wheel; an axial positioning device attached tothe second end of the shifter adapted to move the shifter axially; aposition indicator attached to the steering shaft adapted to indicatethe angular position of the steering shaft; a position detector locatednear the position indicator and that is adapted to detect the angularposition of the steering shaft; and a positioner adapted to axiallyposition the positioning device upon a positioning signal from theposition detector.
 8. The steering system of claim 7, wherein thepositioning device is a lead screw and the positioner is a motorized nutadapted to drive the lead screw.
 9. The steering system of claim 7,wherein the positioning device is a piston and the positioner is acylinder.
 10. The steering system of claim 9 wherein the piston andcylinder are hydraulically operated.
 11. A four wheeled vehicle steeringsystem, comprising four variable speed wheel transmissions, each adaptedto provide torque to one wheel, each of the wheel transmissionscomprising: a longitudinal axis; a plurality of balls distributedradially about the longitudinal axis, each ball having a tiltable axisabout which it rotates; a rotatable input disc positioned adjacent tothe balls and in contact with each of the balls; a rotatable output discpositioned adjacent to the balls opposite the input disc and in contactwith each of the balls; and a rotatable idler coaxial about thelongitudinal axis and positioned radially inward of and in contact witheach of the balls; and a control system adapted to independently controlthe axial position of each of the idlers in response to a request by anoperator and thereby shift a transmission ratio of each of the wheeltransmissions independently such that the wheels of the vehicle can turnat different rates causing the vehicle to turn.
 12. The steering systemof claim 11, wherein each of the wheel transmissions further comprises aplanetary gear set mounted about the longitudinal axis of thetransmission.
 13. A hybrid vehicle, comprising: a first source ofrotational energy; a second source of rotational energy; and atransmission adapted to accept rotational energy from both the first andsecond sources comprising; a longitudinal axis; a plurality of ballsdistributed radially about the longitudinal axis, each ball having atiltable axis about which it rotates; a rotatable input disc positionedadjacent to the balls and in contact with each of the balls; a rotatableoutput disc positioned adjacent to the balls opposite the input disc andin contact with each of the balls; a rotatable idler coaxial about thelongitudinal axis and positioned radially inward of and in contact witheach of the balls; and a rotatable cage adapted to maintain the axialand radial position of each of the balls; wherein the first sourcesupplies rotational energy to the cage and the second energy sourcesupplies rotational energy to the input disc.
 14. The hybrid vehicle ofclaim 13, wherein the first source of rotational energy is an internalcombustion engine and wherein the second source of rotational energy isan electric motor.
 15. The hybrid vehicle of claim 14, furthercomprising an axial force generator adapted to generate a contact forcebetween the input disc, the output disc, the balls and the idler that isproportional to an amount of torque to be transmitted by thetransmission.
 16. The hybrid vehicle of claim 15, wherein the axialforce generator further comprises: a bearing disc coaxial with androtatable about the longitudinal axis having an outer diameter and aninner diameter and having a threaded bore formed in its inner diameter;a plurality of perimeter ramps attached to a first side of the bearingdisc near its outer diameter; a plurality of bearings adapted to engagethe plurality of bearing disc ramps; a plurality of input disc perimeterramps mounted on the input disc on a side opposite of the balls adaptedto engage the bearings; a generally cylindrical screw coaxial with androtatable about the longitudinal axis and having male threads formedalong its outer surface, which male threads are adapted to engage thethreaded bore of the bearing disc; a plurality of central screw rampsattached to an end of the screw facing the speed adjusters; and aplurality of central input disc ramps affixed to the input disc andadapted to engage the plurality of central screw ramps.
 17. A variableplanetary gear set, comprising: a generally tubular idler; a pluralityof balls distributed radially about the idler, each ball having atiltable axis about which it rotates; a rotatable input disc positionedadjacent to the balls and in contact with each of the balls; a rotatableoutput disc positioned adjacent to the balls opposite the input disc andin contact with each of the balls such that each of the balls makesthree-point contact with the input disc, the output disc and the idler;and a rotatable cage adapted to maintain the axial and radial positionof each of the balls, wherein the axes of the balls are oriented by theaxial position of the idler.
 18. The planetary gear set of claim 17,wherein the cage further comprises: an input stator support in thegeneral shape of a disc positioned between the balls and the input disc;an output stator support in the general shape of a disc positionedbetween the balls and the output disc; and a plurality of spacersadapted to extend between and rigidly connect the input stator andoutput stator.
 19. The planetary gear set of claim 18, furthercomprising an axial force generator adapted to provide a contact forcebetween the input disc, the output disc, the balls and the idler that isproportional to the amount of torque to be transferred through the gearset.
 20. The planetary gear set of claim 19, wherein the axial forcegenerator comprises: a generally disc-shaped thrust washer that iscoaxial with the idler and is positioned near the side of the input discfacing away from the balls having a first side facing the input disc andhaving a set of thrust ramps formed on the first side; a set ofthrust-receiving ramps formed on the input disc facing the thrustwasher; and a plurality of thrust elements located between and incontact with the thrust ramps and the thrust-receiving ramps.